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WORKS OF J. A. MOYER 

PUBLISHED BY 

JOHN WILEY & SONS, INC. 



Descriptive Geometry for Students of Engineering 

Third Edition, viii + 204 pages, 6 by 9, 128 figures. 
Cloth, $2.00. 



Steam Turbines 

A Practical and Theoretical Treatise for Engineers 
and Students, including a Discussion of the Gas Tur- 
bine. Third Edition, Revised and Enlarged, xi -j- 468 
pages, 6 by 9, 225 figures. Cloth, $4.00 net. 

Engineering Thermodynamics 

By J. A. Moyer and J. P. Calderwood. 

First Edition, viii + 204 pages, 6 by 9, 71 figures. 

Cloth, $2.00. 



STEAM TURBINES 



A PRACTICAL AND THEORETICAL 

TREATISE FOR ENGINEERS 

AND STUDENTS 

INCLUDING A DISCUSSION OF THE 
GAS TURBINE 



BY 
JAMES AMBROSE MOYER, S.B., A.M. 

MEMBER OF THE AMERICAN SOCIETY OF MECHANICAL ENGINEERS, MITGLIED DES 
VEREINBS DEUTSCHER INGENIEURE, MEMBRE TITULAIRE ASSOCIATION IN- 
TERNATIONALE DU FROID, MEMBER OF THE FRANKLIN INSTI- 
TUTE, AMERICAN INSTITUTE OF ELECTRICAL ENGINEERS, ETC. 

DIRECTOR OF THE MASSACHUSETTS DEPARTMENT OF UNIVERSITY EXTENSION, 
FORMERLY PROFESSOR OF MECHANICAL ENGINEERING IN THE PENN- 
SYLVANIA STATE COLLEGE, ENGINEER, WESTINGHOUSE, 
CHURCH, KERR & CO., AND STEAM TURBINE DEPART- 
MENT, GENERAL ELECTRIC COMPANY 



THIRD EDITION, REVISED AND ENLARGED 
FIRST THOUSAND 



NEW YORK 

JOHN WILEY & SONS, Inc. 
London: CHAPMAN & HALL, Limited 

1917 



<■ 



Copyright, 1908, 1914, 1917 



JAMES AMBROSE MOYER 




b 







u 



FEB -8 1917 

Stanbopc ipress 

F. H.GILSON COMPANY 
BOSTON, U.S.A. 



©CU455480 



PREFACE TO THIRD EDITION 



Epoch-making developments in the production of power are 
apparent in the recent and general application of very large steam 
turbines. The remarkable economy of large turbines, com- 
pared with all other prime movers using steam, is causing the 
replacement of small engines by larger and larger turbines. 
Concentration in power plants of tremendous capacity of these 
large units is revolutionizing industry. Unlimited use of elec- 
tricity seems to be a certainty of the near future which must 
be obtained largely from huge central power stations operated 
with very large steam boilers and turbines with their direct- 
connected electrical generators. With this development of 
steam power stations, the so-called " isolated" steam plants can 
only be justified when practically all the exhaust steam can be 
used during a considerable part of the year for heating purposes. 

Recent improvements in the economy of steam turbines are 
having also an important effect on general industry. The low 
cost of power where coal is cheap makes the large turbine-electric 
generating plant almost an unrivaled competitor of water power 
for metallurgical processes, such as the reduction of aluminum 
from its ores, and the refining of metals such as copper. 

In spite of the general increase in the cost of raw and manu- 
factured materials in the last few years, the application of steam 
turbines in power plants to take the place of reciprocating 
engines has reduced the total first cost of large first-class power 
plants from $120 per kilowatt of rated capacity, a fair aver- 
age value of five years ago, : to the present cost of about $60. 
Electro-chemical industries are also being benefited by the avail- 
ability of this cheap steam power for the fixation of the nitro- 
gen in air for use in fertilizers. Finally, the electric propulsion 

iii 



IV PREFACE 

of ships, now receiving so much attention from engineers, and 
the electrification of the eastern trunk-line railroads will be a 
rapid development with the assistance of cheap power from 
large steam turbine generating plants. 

Steam turbine units to generate 70,000 kilowatts are now a 
reality, with an economy of only a little more than ten pounds 
of steam per kilowatt-hour. Since a man can take care of a 
large unit almost as easily as a small one, the cost of attendance 
per unit of power goes down like a feverish " break" in the 
stock market, in proportion to the size of the unit. Fifteen 
years ago the best American power plants could, under the 
most favorable conditions, do no better than develop a kilowatt- 
hour from three pounds of coal. Today it is easily possible to 
produce the same power from half the amount of coal. These 
are matters of great significance to industry. 

In these times when there is such general discussion of con- 
servation and efficiency, the low-pressure steam turbine takes 
an important place, because of its innumerable applications for 
preventing the wasting of any steam to the atmosphere. In 
nearly every large central station hundreds of pounds of steam 
exhausted from the auxiliaries at atmospheric pressure in excess 
of that required in the heaters for heating the feed is lost through 
the exhaust-heads. Modern methods of turbine application 
would save and utilize this steam for power. In the most 
modern practice, therefore, the greatest skill of the engineer is 
called upon more in connection with the methods of applying 
the commercial types of turbines already developed rather than 
in the actual designing of new types of machines. In steam 
engine designing there have been always unlimited possibilities; 
in modern steam turbine designing these are few. 

Most of the additions made in this edition have been, there- 
fore, mainly in the line of new applications. The chapter on 
low-pressure turbines has been rewritten and very much ex- 
tended to include the latest developments and applications. 
This chapter should be unusually interesting to all engineers 
and students. The chapter on Reaction Turbine Design has 



PREFACE V 

been entirely rewritten with special reference to the problems 
arising in the design of combined impulse and reaction types. 

Many changes have been made in practically all the chapters 
of the book, embodying, in many cases, the suggestions of pro- 
fessors in colleges and universities where the book is used as 
a text. Many additions have been made in the chapter on Com- 
mercial Types to make the descriptions in every sense up-to-date. 

In the preparation of this edition I am particularly indebted 
to Professor J. E. Emswiler of the University of Michigan, 
Professor J. G. Callan of the University of Wisconsin, Professor 
E. A. Fessenden of The Pennsylvania State College, and Mr. M. 
Nusim of the Southwark Foundry and Machine Company, 
Philadelphia. 

Acknowledgment should be made of important criticisms 
and suggestions received from Professor J. B. Ludy of Purdue 
University, Professor W. E. Brooks of the University of Min- 
nesota, Dr. William Kent of Montclair, New Jersey, Professor 
J. J. Eames of Massachusetts Institute of Technology, Mr. F. 
A. Burr of Providence, Rhode Island, and Mr. A. R. Meek of 
Boston, Mass. 

Papers prepared by Messrs. H. T. Herr, Vice President and 
General Manager of the Westinghouse Machine Company, 
Richard H. Rice, Consulting Engineer of the General Electric 
Company, and H. G. Stott, Superintendent of Motive Power, 
Interborough Rapid Transit Company of New York have been 
freely consulted in this revision. The bulletins published by 
Allis- Chalmers Company of Milwaukee, and De Laval Steam 
Turbine Company of Trenton, New Jersey, have been very 
useful and a number of illustrations have been taken from them. 

THE AUTHOR. 

State House, Boston, Mass. 
January 4, 191 7. 



PREFACE TO FIRST EDITION 



The object of this book is to give in a small volume what I 
believe, as the result of years of practical experience, engineers 
and students of engineering want to know about steam turbines. 
It is intended that it shall be a manual for the practical engineer 
who is designing, operating, or manufacturing steam turbines 
rather than a compilation of manufacturers' catalogs combined 
with a digest of standard books on thermodynamics and 
mechanics. 

In a general way the author has tried to explain briefly and 
directly some of the more important problems about which the 
qualified steam engineer must have some knowledge. When this 
book was first planned it was intended primarily for the use of the 
author's assistants in the experimental and testing departments 
of one of the large manufacturing companies, but later it seemed 
that it might be useful in a larger field. 

The order in developing the subject is the reverse of that adopted 
by most authors. Instead of discussing the intricacies of blading 
in the beginning of the book, the more simple problems of nozzle 
design are presented first. A great deal more is now known about 
nozzles than there was even very few years ago, and many of the 
conditions affecting the efficiency of nozzles may now be considered 
well established. Nozzles are also becoming a more important 
part of all types of turbines. Even the Parsons turbine is now 
being modified in America and England so that in many of the 
latest designs for large sizes, nozzles are used in the high-pressure 
stages. It is coming to be generally recognized that in the future 
there will probably be no large ' installations of reciprocating 
engines for electric services. A few years ago this might have 
been considered a bold statement, but it is a fact which is now 



vui PREFACE 

generally, although reluctantly, admitted by manufacturers of 
reciprocating engines. 

The entropy-total heat chart in the back of the book is laid out 
with lines of constant superheat instead of lines of constant tem- 
perature which have been generally used for charts of this kind. 
For practical engineering work it is very desirable to have lines 
of constant superheat on such charts, because in America and 
England guarantees of steam consumption are usually given in 
degrees of superheat rather than of temperature. When charts 
made with constant-temperature lines are used, it is always neces- 
sary to calculate the temperature before the chart can be used. 

Most of the graduates of our American technical schools are 
entirely "at sea" with the simplest heat calculations, and one of 
the reasons for this deficiency is that most of the books on steam 
engines — and especially those on the steam turbine — are more 
devoted to giving a large quantity of facts than to fulfilling a useful 
purpose. Practical engineers who have had to deal with large 
numbers of men with an engineering training agree most candidly 
with Dr. Steinmetz when he says in substance that it seems to cause 
no concern in some of our large technical schools that the graduates 
are sent out loaded with a mass of half-understood and undigested 
subjects, while they are deficient both in the understanding of the 
fundamental principles and in the ability to think. If this volume 
can serve the purpose of encouraging students to think it will 
have accomplished one of its principal purposes, not losing sight 
of the fact that the book is intended primarily to show how to 
do things. 

Nearly all the proof-reading has been done by Professor John F. 
Pelly of Philadelphia. Because of Professor Pelly's thoroughly 
practical as well as theoretical knowledge of the subject matter, his 
conscientious and painstaking work is very greatly appreciated. 

I take this opportunity to thank Professor Ira N. Hollis and 
Professor F. Lowell Kennedy of Cambridge for the criticisms and 
suggestions which I received from them when the manuscript of 
this book was preparing. I am also greatly indebted to Mr. Walter 
C. Kerr, president, and Mr. Sidney E. Junkins, vice-president of 



PREFACE ix 

Westinghouse, Church, Kerr & Company, for their encouragement 
and for making it possible to finish the book at this time. 

For placing at my disposal a great deal of information regarding 
the latest results in steam turbine engineering, which is usually 
very difficult to obtain, I am particularly indebted to Mr. Richard 
H. Rice of Lynn, and Mr. J. R. Bibbins of Pittsburg. 

I wish to thank Professor Arthur M. Greene of Troy and 
Mr. Albert Stritmatter of Cincinnati for suggestions relating to the 
subject matter. For various services in the preparation of this 
book, I should mention also Messrs. Francis Hodgkinson and 
Harold P. Childs of the Westinghouse Machine Company; C. P. 
Crissey, S. A. Moss, and W. E. Culbertson of the General Electric 
Company; C. P. Chasteney of the De Laval Steam Turbine 
Company; James Wilkinson, president of the Wilkinson Turbine 
Company; St. John Chilton of the Allis-Chalmers Company; H. H. 
Wait of the Western Electric Company; Carl S. Dow of the B. F. 
Sturtevant Company; and J. Clarence Mover of Philadelphia. 

Many of the illustrations for the book have been provided, in 
some cases at considerable expense to themselves, by the Cassier 
Magazine Company, Westinghouse Machine Company, General 
Electric Company, De Laval Steam Turbine Company, Rateau 
Turbine Company, Kerr Turbine Company, Wilkinson Turbine 
Company, Allis-Chalmers Company, C. H. Parsons & Co., and 
Brown, Boveri & Co. 

Throughout the text important words and sentences are brought 
out by the use of bold-faced type, thus making the subjects of a 
paragraph visible at a glance. 

The author is always glad to answer correspondence with 
teachers relating to questions which inevitably arise in the discussion 
of designs for steam turbines, all of which cannot, of course, be 
taken up in detail in any book. 

JAMES AMBROSE MOYER. 

New York City, 
September, 1908. 



CONTENTS 



CHAPTER PAGE 

I. Introduction i 

II. The Elementary Theory of Heat 10 

III. Nozzle Design 36 

IV. Steam Turbine Types and Blade Design 60 

V. Mechanical Losses ln Turbines 151 

VI. Method for Correcting Steam Turbine Tests 162 

VII. Commercial Types of Turbines 176 

De Laval, Parsons, Westinghouse, Allis-Chalmers, Curtis, Rateau, 
Wilkinson, Zoelly, Sturtevant, Riedler-Stumpf, Kerr, Terry, 
Dake, etc. 

VTII. Governing Steam Turbines 274 

LX. Low-Pressure (exhaust) Turbines 311 

X. Mixed-Pressure Turbines 335 

XI. Bleeder or Extraction Turbines 339 

XLT. Marine Turblnes 345 

XIII. Tests of Turbines 353 

XTV. Steam Turblne Economics 36$ 

XV. Stresses ln Rings, Drums, and Disks 405 

Design of Turbine Wheels — Critical Speeds of Loaded Shafts. 

XVT. Gas Turbines 432 

XVII. Electric Generators for Turbines 450 

Direct-current Generators — Alternating-current Generators. 

Appendix of Practical Exercises 455 



THE STEAM TURBINE 



CHAPTER I. 
INTRODUCTION. 

The steam turbine is the most modern as well as the most 
ancient steam motor. Recently its development has gone by 
leaps and bounds; and, above all, in its applications it is gaining 
ground daily. Doubtless it is to be the most important prime 
mover of the near future. 

During recent years results have been secured with steam tur- 
bines that only a short time ago were considered practically unat- 
tainable. Primarily their great success lies in their adaptability 
to operation with high vacuums. Steam turbines are, therefore, 
almost ideally suitable for the conditions of modern engineering 
practice requiring both high vacuums* and high superheats. 
To-day in the economical use of steam they are unrivaled; and, 
because of improved manufacturing methods, marking the tran- 
sition from the experimental to the commercial stage, first cost 
is no longer a deciding factor favoring reciprocating engines. 

Compared with reciprocating steam and gas engines, steam 
turbines require much smaller and cheaper foundations, occupy 
less floor space, require fewer attendants, and because no lubrica- 
tion is required for any parts in contact with the steam, the con- 
densation becomes directly available for feed water. The highest 
superheats can be employed without affecting the choice of lubri- 
cants, and the cost of oil for lubrication is very low. 

A steam turbine of the simplest type is essentially a wheel 
similar to an ordinary water wheel, which is moved around by 
a steam jet impinging on its blades. Steam is directed against 
the turbine wheel by nozzles or similar passages delivering the 

* The question of the most profitable vacuum for given conditions is discussed 
on pages 365 to 373. 



THE STEAM TURBINE 




Fig. i. A Small Modern Steam Turbine with Part of the Casting Removed. 

steam at mathematically exact angles, calculated to make the 
steam strike the blades of the wheel most advantageously. 

Fig. i is an illustration of a modern steam turbine with a part 
of the casing removed to show the construction. The turbine 



INTRODUCTION 



wheel W is shown here with numerous blades on its circumference. 
The steam comes to the turbine from the boilers through a suitable 
steam main connected to the top of the "turbine at M and passes 
down through the pipe A to the steam-chest B. From this steam- 




Fig. 2. The Turbine Wheel and Nozzles. 

chest it is guided through one or more nozzles, from which it 
escapes at a high velocity to impinge on the blades on the circum- 
ference of the turbine wheel, which is thus made to rotate, and 
performs work by moving machinery connected to the shaft. 
Nozzles from which the steam is discharged are located around 
the periphery of the wheel as shown in Fig. 2 with their enlarged 



THE STEAM TURBINE 



ends, technically called mouths, very close to the blades.* Steam 
after passing through the blades enters the exhaust pipe at E 
(Fig. i ) and is discharged into the atmosphere or into a condenser, 
depending on whether the operation is non-condensing or con- 
densing. 

Preliminary to the study of the modern commercial types of 
steam turbines it is desirable to state briefly some of the most 
important stages through which this very ancient form of steam 
motor has passed in its development. 

Early History. The earliest notices of heat engines of any 
kind are found in a book by Hero of Alexandria, which was 
probably written in the second century before Christ. In this 
book of mechanical contrivances a steam reaction wheel is men- 
tioned. This first steam turbine is shown 
in Fig. 3. It is described as consisting 
of a hollow spherical vessel pivoted 
on a central axis and supplied with 
steam through the support M and one 
of the pivots from a boiler, B, beneath. 
Steam escaped from the vessel through 
bent pipes or nozzles N, N, facing tan- 
gen tially in opposite directions. The 
spherical vessel was revolved by the re- 
action due to the escaping steam, just as 
a " Barker's mill " is moved by the water 




Fig. 3. Hero's Turbine. 



escaping from its arms. Any fluid escap- 
ing under pressure from a vessel which 
is free to move causes a "reaction" tending to displace the vessel 
in the opposite direction from the flow of the fluid. This reaction, 
although imperfectly understood by Hero, was perfectly applied 
in his steam turbine which was used to open the doors of 
temples. Only a few years ago a model of Hero's engine was 
constructed by a celebrated English engineer,! with, of course, 

* In this figure one of the nozzles is represented as if transparent to show its 
shape on the inside, and a part of the steel band around the blades is cut away to 
show the shape of the blades or vanes, as well as to illustrate the passage of steam 
from the nozzle into and through the blades. 

t See page 8. 



INTRODUCTION 



all the advantages of modern machine tools and appliances, 
with the result that an engine was produced which, in economy, 
compared well with our elaborate and complicated modern 
engines. 

In 1577 a German mechanician, it is said, used a turbine simi- 
lar to Hero's to rotate reaming and burnishing tools, but from 
the time of Hero down to the seventeenth century there is no 
record of progress in the development of steam heat engines. In 
1629 Branca, an Italian architect, designed a steam turbine 
(Fig. 4) resembling a water wheel, which was driven by the 
impulse from a jet of steam directed by means of a nozzle upon 
suitable vanes attached to the 
wheel. Branca' s turbine en- 
gine, however, was not success- 
ful; and until the end of the 
nineteenth century, although in 
the interval many steam tur- 
bines and other rotary engines 
were patented, the piston or 




reciprocating steam engine, 



Fig. 4. Branca's Turbine. 



under the leadership of Watt, 

had, commercially, an unrestricted field and remarkable results 

were accomplished. 

It is interesting to observe that the modern type of impulse 
turbine with a single row of blades like the one illustrated in 
Fig. 1 is practically the same, except for details, as the historic 
Branca's wheel. The principal difference is that Branca's wheel 
was not enclosed in a casing. Essential parts — the nozzle, 
the blades, the wheel, and the shaft — were practically the same 
as in some modern machines. Probably if Branca had under- 
stood the laws of the expansion of steam as we do to-day, he could 
have made a successful prime mover of his turbine. Those who 
came after him were aided not only by a superior knowledge 
but also by the opportunities for scientific investigation and the 
skill of our present-day workshops. 

De Laval Type. Dr. Gustaf De Laval, a Swedish scientist, was 



6 THE STEAM TURBINE 

a pioneer in the modern commercial development of steam tur- 
bines. In 1882 he constructed his first steam turbine, which was 
similar in principle to Hero's reaction engine. De Laval's first 
turbine was designed primarily for driving his milk and cream 
separators, for which there was then a large sale. For other 
purposes, however, there was no general application, because at 
the very high speeds for which they were designed, it was difficult 
to utilize the power; and besides, the steam consumption was 
practically prohibitive. 

Later De Laval turned his attention to the development of 
Branca's steam turbine, and was remarkably successful. After 
much experimenting, he developed an impulse turbine which is 
still one of the standard makes. (See Figs. 82 to 86.) This 
great engineer, after investigating the possibilities of both Hero's 
and Branca's types and having decided to adopt the latter, began 
then some strikingly original inventive work, which, in many 
respects, led the way for the accomplishments of to-day. 

It should be stated, however, in this connection, that no engineer 
thinks of belittling De Laval's work because his investigations 
were mostly in the line of improvements to existing types. Un- 
questionably he must have the credit for producing the first 
commercially successful steam turbine. Many of the features of 
his original designs have actually contributed in no small meas- 
ure to our knowledge of machine design and thermodynamics, 
and have become fundamental principles underlying many of the 
most important modern steam turbine developments.* 

Parsons Type. With the early work of De Laval, however, 
the development of steam turbines designed to operate by the 
reaction principle of Hero's engine was not given up. Almost 
contemporaneously with De Laval, C. A. Parsons in England 
began the development of the well-known type which to-day bears 
his name, and which has made possible the brilliant records of 
turbine ocean steamers. In April, 1884, this great inventor took 

* The most important feature introduced by De Laval is that of the diverging 
nozzle (British Patent No. 7143 of 1889), the principle of which has influenced 
the development of practicallv all types of steam turbines. 



INTRODUCTION 7 

out his first patents on steam turbines. The practicability of the 
steam turbine he then proposed is a striking feature of even his 
first patents. His specifications showed, above all, that a great 
deal of time and thought had been devoted to constructive details. 
Methods for reducing vibration, preventing leakage of steam, and 
providing for efficient lubrication contributed very largely to his 
success. Many of the details of this early turbine are now obso- 
lete, so that only a very short description will be given here. A 
section drawing of Parsons' first turbine is shown in Fig. 5. A 




Fig. 5. Early Parsons Steam Turbine. 

large central collar, C, is attached to the main shaft, S, which runs 
the length of the turbine. At the ends of the casing where the 
shaft passes through it the cross-section is reduced. The main 
shaft, S, supports a large number of rings which are held in place 
between the collar, C, and the nuts, N, which are screwed on the 
reduced section of the shaft at the ends. These rings, around 
their circumferences, support those turbine blades (b, b) which 
move with the shaft. There are, however, alternating with them, 
other rows of blades (c, c) attached to the inside of the turbine 
casing. Technically the blades b, b are called moving blades, 
and c, c are called fixed or stationary blades. Steam is admitted 
to the turbine blades through the annular chamber, A, encircling 
the collar, C, and then it passes to the right and left through the 
alternate rows of stationary and moving blades to the exhaust 
passages E, E — one at each end of the turbine. The steam 
expands in the blades as in a nozzle, and its reaction moves the 
blades attached to the shaft, just as Hero's turbine was rotated 
by the steam escaping from its arms. 



8 



THE STEAM TURBINE 



By the " double- flow " arrangement in this design by which the 
steam is passed from the center to the exhaust at both ends there 
can be very little axial thrust on the shaft. Any thrust that 
does occur, however, is balanced by the pressure of the exhaust 
steam in the chambers E, E at the ends of the casing. A slight 
movement of the shaft toward either end checks the flow of the 
exhaust steam and increases the back pressure at that end. This 
increased pressure then moves the shaft back to its normal 
position. 

Usually it is not possible to balance the parts of a rotating mass 
to make its center of gravity coincide exactly with the geometric 
center about which it revolves. In any machine like a steam 
turbine, when these two centers do not coincide excessive vibra- 
tions of the shaft are produced which at certain speeds * are 
sufficient to break it. To overcome this difficulty, Parsons in- 
geniously allowed a little lateral play, or " elasticity," as he called 




Fig. 6. Screw Type of Steam Turbine. 



it, for the shaft by means of a series of rings of two different 
diameters, in principle very much the same as the present con- 
struction of the main bearings of Parsons turbines (see Fig. ioo), 
so that it was permitted to move laterally a certain amount, say a 
hundredth of an inch, to allow the proper adjustment in passing 
from rest to the normal speed of running. 

Among his early experiments Parsons also tried a purely 
reaction steam turbine, following almost exactly the published 
designs of Hero. This turbine, running with ioo pounds per 

* This phenomenon occurs at very definite speeds, called "critical," for every 
rotating mass. Fuller discussion, with a method for calculating "critical" speeds, 
is given on page 430. 



INTRODUCTION 9 

square inch steam pressure and 27 inches of vacuum, gave an 
output of 20 horsepower at 5,000 revolutions per minute. Steam 
consumption was only 40 pounds per brake horsepower per 
hour, which was indeed a remarkably good result for that time. 

Screw Type. Still another kind of turbine, of only historical 
interest, should be mentioned. A large number of inventors have 
worked on the development of a screw type like Fig. 6. Hewitt 
worked for a long time on a turbine of this kind, and finally con- 
cluded the results were not satisfactory. Steam was admitted to 
this turbine through the chamber A, and passed through holes in 
the plates P, P into the helical grooves on the shaft. In these 
grooves the steam was expanded and then escaped to the exhaust 
pipes E, E at the two ends. Effective action of the steam was 
probably obtained only in the first part of the grooves; and after 
being deflected into a helical course, it rushed through to the 
exhaust without much additional effect in moving the shaft. 
Excessive leakage of steam between the helical threads and the 
casing was another serious difficulty. 

Recently a somewhat similar arrangement having two " screw 
wheels " meshing together not unlike spiral or helical gears has 
been successfully developed by the Buffalo Forge Company (see 
Fig. 157, page 270). 



CHAPTER II. 

THE ELEMENTARY THEORY OF HEAT. 

Note. — This short chapter may well be omitted, in reading, by those who are 
familiar with the thermodynamics of heat engines and with the use of entropy 
diagrams. It is intended primarily for practical engineers, who will find it par- 
ticularly valuable for reference purposes, as the subject matter is completely 
indexed. 

Technically the steam turbine must be regarded as a heat 
engine, that is, a machine in which heat is employed to do mechan- 
ical work. From the viewpoint of the practical man its function, 
the same as that of any other heat motor, is to secure as much 
work as possible from a given amount of steam, or, going a step 
farther back, from the combustion of a given amount of fuel. 
Heat theory is, therefore, of first importance. 

Heat is a form of energy like electrical, chemical, mechanical, 
potential, and kinetic. No doubt exists about the equivalence 
of the different forms of energy and their close relation to each 
other. Each, at will, can be changed into any of the other forms. 

The relative amount of heat in a body is observed, in common 
experience, by the sense of touch — whether the body is a solid, 
a liquid, or a gas. By such experience we have learned to recog- 
nize certain sensations as hot or cold; and then, with more 
accuracy, to speak of degrees of temperature. Now when a hot 
and a cold body are brought together their temperatures become 
equalized. The hotter body always loses heat. The colder 
body always gains heat.* This experience is the principal basis 
for all heat calculations. 

When in the course of time it had been found that a more 
accurate method than that of the sense of touch was needed for 
heat determinations, methods utilizing the expansion of liquids 

* This phenomenon is called the second law of thermodynamics, — that "heat 
energy always passes from a warm body to a cold body." 

10 



THE ELEMENTARY THEORY OF HEAT II 

came to be generally employed. Many substances have a practi- 
cally uniform rate of expansion between the limits of temperature 
an engineer has to deal with. A small column of mercury in a 
glass capillary tube is usually taken as a standard for temperature 
measurements.* The mercury in an accurate thermometer 
expands very nearly ^|^ of its volume when heated from the 
freezing temperature of water (32°F.) to the boiling point 
(2i2°F.). The expansion between the freezing point and the 
boiling point of water has therefore been called, arbitrarily, 
180 F. 

For theoretical heat calculations the zero of temperature is 
taken as 492 F. below the freezing temperature of water; or, 
460 below the Fahrenheit zero. This very low temperature is 
called the absolute zero, and at this point there is theoretically 
no heat energy. 

Temperatures measured from the absolute zero are called 
absolute temperatures and are indicated generally by T, to dis- 
tinguish them from the ordinary Fahrenheit temperatures, t, as 
read on a thermometer scale. 

Using these symbols, we have then in Fahrenheit degrees, 

T = t + 460. 

Absolute temperatures are convenient for heat calculations 
because " perfect " gases, at constant pressure, increase in volume 
in proportion to the increase in absolute temperature. 

* The ordinary mercury thermometers can be used to measure temperatures to 
about 575 F. with accuracy. For higher temperatures the capillary tube over 
the mercury should be filled with nitrogen or carbonic acid gas under high pressure. 
Such thermometers can then be used for temperatures up to iooo° F. 

If the mercury is not throughout its whole length at the same temperature as 
that being measured, a correction, k, given by the following formula must be 
added to the observed temperature, /, in Fahrenheit degrees: 

k = .ooo,o88 2? (*-/'), 
where D is the length of the mercury column exposed, measured in Fahrenheit 
degrees, and t' is the temperature of the exposed part of the thermometer. When 
long thermometers are used in shallow wells in high-pressure steam pipes this cor- 
rection is often 5 to io° F. For experimental data and direct-reading correction 
curves, see Mover's Power Plant Testing, 2d edition (McGraw-Hill Book Co.), 
pages 31-33. 



12 THE STEAM TURBINE 

Heat Units and Specific Heat. The amount of heat required 
to raise the temperature of one pound of water from 62 to 63 F. 
is taken arbitrarily as the standard English unit of heat, — com- 
monly called the British thermal unit (B.T.U.).* The ratio of 
the amount of heat required to raise the temperature of a pound 
of water or steam one degree to the British thermal unit is called 
the specific heat.f 

The specific heat of steam and of gases changes in value accord- 
ing to the conditions under which the heat is applied. If heat is 
added to a vapor or a gas held in a closed vessel, with no chance 
for expansion, no external work is done, and therefore practically 
all the heat added is used to increase the temperature. This is 
the condition in a boiler when no steam is being drawn off. In 
this case the specific heat is symbolized by C v = specific heat at 
constant volume. If, on the other hand, the pressure is kept 
constant but the volume is allowed to change to permit expansion 
and the performing of external work, we say then, C p = specific 
heat at constant pressure. 

Heating at constant pressure is the condition that is most 
interesting to the engineer. When his engines are running the 
boilers are making steam at constant pressure. The heat energy 
absorbed by a pound of steam for raising only the temperature 
must be, obviously, approximately the same, regardless of the 
conditions of pressure and volume. Since for constant pressure 
conditions some external work is always done, requiring a 
larger amount of heat energy than for the case when the volume 
is constant, it follows that C p is always greater than C v . 

We should add, further, that an engineer's calculations con- 
cerning energy transformations in steam turbines are almost 

* In the C. G. S. system of units the kilogram-calorie, called in German 
Warmeeinheit (WE), is used as the standard heat unit. 1 kg.-cal. or 1 WE = 3.97 
or nearly 4 B.T.U. 

f The specific heat of water at 200 F. is 1.005, an d of superheated steam an 
average value of .6 is often assumed in rough calculations for steam at the usual boiler 
pressures in power plant practice for superheats less than 150 F. Mean values of 
the specific heat of superheated steam are given by the curves in Fig. 12. 



THE ELEMENTARY THEORY OF HEAT 13 

without exception for the condition of constant pressure, and, 
consequently, only values of C p are generally useful. Most gases 
have practically constant values for their specific heats. 

At temperatures near the boiling point, the heating of vapors, 
like steam, is influenced by molecular attraction, so that their 
specific heats are variables depending on conditions of tempera- 
ture and pressure. The specific heat of superheated steam de- 
creases with increasing temperatures to a minimum value. The 
values of specific heat increase slightly, on the other hand, with an 
increase of pressure.* 

Mechanical Equivalent of Heat. Heat and work are both 
forms of energy and are " equivalent," meaning that energy can 
be transformed into mechanical work, and that work, as a form 
of energy, can be changed back again into heat. The relation is 
expressed by 

1 British thermal (heat) unit = 778 foot-pounds (work). 

HEAT AND WORK. 

Heat is a form of energy. Each of the various kinds of heat 
motors, such as the steam engine, the steam turbine, the gas 
engine, or the gas turbine, is a machine for obtaining mechanical 
work from heat energy. 

In the general principles of operation the steam turbine and the 
reciprocating or " piston " steam engine are essentially similar 
machines. Both do work according to the same heat relations. 
The gas turbine is somewhat different. This new motor, which 
as yet has scarcely reached a practical stage of development, 
will be discussed in its proper place. 

In a reciprocating steam engine working " expansively " the 
steam is admitted at boiler pressure until the point of cut-off; and 
during the remainder of the stroke the piston is pushed ahead, or 
does work, by the expansion of the steam shut up in the cylinder. 
In the steam turbine the heat process is analogous, except that 

* Knoblauch and Jakob, Zeit. Verein deutscher Ingenieure, Jan. 5, 1907, and 
an article by the author in Mechanical Engineer (London), Aug. 24, 1907. 



14 THE STEAM TURBINE 

the flow of the steam, instead of being intermittent, is continuous. 

Steam is continually pushed into the nozzles, or similar steam 
passages, and expanding, expends its internal energy in producing 
velocity. Vanes or blades, fixed to a rotating wheel, are placed 
near the nozzles so that the jets of steam are directed against 
them. These blades or vanes thus set in motion move the wheel 
and with it the shaft which transmits the power. 

Theoretically the work from expanding steam behind a piston 
is exactly the same as that we obtain from a nozzle. The difference 
is only in the method for making the heat energy available for 
doing work. 

Before going farther with the discussion of how the steam tur- 
bine converts heat energy into work, the more familiar case of the 
reciprocating steam engine will be considered briefly, because it 
is assumed the reader is already more or less familiar with its 
heat processes. By the static pressure in the steam pipes and in 
the boiler the steam is pushed into the engine cylinder and causes 
the piston to move up to the point where the supply of steam is 
shut off. Then the steam expands, reducing, at the same time, 
the pressure till the piston has reached the end of the cylinder. 
On the return stroke the steam is discharged at a nearly constant 
low pressure into the atmosphere or into a condenser. Now on 
the " working " stroke when the steam is being pushed into the 
cylinder,* and when it is expanding, the steam is doing work at 
the expense of the heat energy put into it by the fires under the 
boiler. The heat in a pound of steam at a given pressure and 
temperature represents a definite amount of energy. Expansion 
of the steam in the cylinder after cut-off is accompanied, therefore, 
with a reduction of pressure and temperature, and the work done 
is in proportion to the heat energy lost by the steam. Thus heat 
energy and work go hand in hand. A loss to one is a gain to the 

* Until the point of cut-off is reached, all the time that steam is being pushed 
into the cylinder work is being done at the expense of the boiler pressure. Actually 
the pressure in the boiler is a little lower after the amount of steam required for a 
stroke has been taken out than it was before. When, however, the strokes of 
the engine come in quick succession, the variation in boiler pressure is not per- 
ceptible. 



THE ELEMENTARY THEORY OF HEAT 



15 



other. Fig. 7 shows a typical steam engine indicator card, repre- 
senting, diagrammatically, the heat relations that have just been 
discussed. The horizontal scale of coordinates (abscissas) repre- 
sents volumes, and the vertical scale (ordinates) represents 
pressures. It is obvious then that any area included by the lines 
of this diagram represents work done by the steam. In this 
figure P t and v t represent initial pressure and volume, and P 2 and 
v 2 the corresponding final conditions, meaning the pressure and 
volume at the end of the "working" stroke. This diagram as 
it applies to the steam engine may be analyzed briefly as follows: 




Fig. 7. Pressure-Volume Diagram Showing Work Areas. 

i. Area AOiB is the work done in " pushing" the steam into 
the cylinder against the resistance of the piston to motion. 

2. Area 12CB is the work done when the steam is expanding. 

3. Area A432C is the work lost in the heat energy discharged 
in the exhaust.* 

4. Area 40123 is the net work done. 

The discussion given above is, of course, for the ideal case 
where the cylinder clearance is neglected and expansion to 
back-pressure (P 2 ) is complete. 

* If an almost perfect vacuum were attainable this loss would be practically 
negligible. Actually with the best condensing apparatus it is quite large. 



l6 THE STEAM TURBINE 

The same diagram (Fig. 7) can also be used for the analysis of 
the work done by steam expanding in the nozzles or similar 
passages * of a turbine. The work done in " pushing " the steam 
into the engine cylinder has its counterpart now in the work done 
by the steam in entering the nozzle, so that, 

1. Area AOiB is the work done in " pushing " the steam out of 
the pipes or receiving vessels into the nozzle. f 

2. Area 12CB is the work done during expansion at the expense 
of the heat energy, to give velocity to the steam. 

3. Area A432C is work lost by the steam in forcing its way 
against the external or exhaust pressure. 

4. Area 40123 is the work done in producing velocity. 

The work of " pushing " the steam into the nozzle produces 
initial velocity!, or "velocity of approach." In all practical steam 
turbine nozzles this initial velocity, compared with the final 
velocity after expansion, is very small. For this reason, in the 
calculations required for the designing of nozzles and blades, 
this initial velocity is usually neglected. Practical designers, 
therefore, are interested only in the heat energy of the area 123 
and the velocity it represents. In order to secure high efficiency 
and low steam consumption the designer is always striving to 
make this area as large as possible, allowing, of course, for other 
limiting conditions. 

As the result of the comparison of the heat functions of steam 
turbines and reciprocating steam engines, we should observe, 
then, that the heat energy in a pound of steam available for per- 
forming useful work is exactly the same whether the steam goes 
to the one or to the other. It follows then also that, theoretically, 

* In some types of turbines there are no nozzles, but instead stationary blades 
are used which are arranged to expand the steam just as in a nozzle. In this 
chapter, therefore, where the term "nozzle" is used it will be assumed to apply as 
well to stationary "expanding" blades. 

t The amount of this work, or the area AOiB, is very small in the case of the 
turbine compared with that in the steam engine. 

V 2 

t This initial velocity, V , is calculated from the relation P x v x = , where 

P x and v x are the initial pressure and volume of a pound of steam and g is the 
acceleration due to gravity (32.2). All velocities are in feet per second. 



THE ELEMENTARY THEORY OF HEAT l? 

the steam consumption for the same conditions of temperature and 
pressure is the same for the turbine as for any other form of 
engine. Discussion of the merits of different forms of steam 
motors with only the theoretical viewpoint in mind is, therefore, 
useless. Only the conditions in practice affecting the design of 
commercial machines are of any significance in determining the 
type of steam motor to be used for given conditions of service. 

HEAT THEORY RELATING TO THE DESIGN OF NOZZLES 
AND BLADES. 

Diagrams similar to those made on a steam engine indicator 
(Fig. 7), showing for an engine stroke the conditions of pressure 
and volume inside the cylinder, are very useful in the design and 
operation of reciprocating steam engines, but they are of very 
little use for work relating to steam turbines. In a steam tur- 
bine it is not practicable to put a measured amount of steam 
through a nozzle " at a time " as the flow is practically con- 
tinuous. The pressure-volume diagram has, therefore, a very 
limited application. Another kind of diagram, the details of 
which are somewhat more difficult to understand, is universally 
used by steam turbine engineers. In this diagram, which will 
now be described, any surface represents accurately to given 
scales a quantity of heat. Absolute temperatures (T) are the 
ordinates, and entropies * (</>) are the abscissas. 

* Entropy, which Perry calls the " ghostly quantity," has no real physical 
significance, so that complete definition is not possible. If dQ is a small amount 
of heat added to a body, and T is the absolute temperature at which the heat is 

added, then the change in entropy of that body is y , or ^ = y ■ Entropy of 

saturated steam above the entropy of water at the freezing point (3 2° F.) is easily 
calculated; but for low temperatures, its values are unknown because of the lack of 
data regarding the specific heats. For saturated steam at any pressure, then, 

$ = ^ + n (or 0), where x is the quality of the steam, r is the latent heat of 

evaporation or " heat of vaporization," T is the absolute temperature, and n (or 6) 
is the entropy of the liquid (water). All values of latent heat of evaporation, heat 
of the liquid, total heat, etc., given in steam tables are in heat units above 32 F. 

The symbols used here are those given in Peabody's Steam and Entropy Tables, 
published by John Wiley & Sons, New York, and in Marks and Davis' Steam Tables 
and Diagrams, published by Longmans, Green & Co., New York. 



i8 



THE STEAM TURBINE 




Fig. 8. 



A Simple Temperature-Entropy 
Diagram. 



Fig. 8 shows a simple heat diagram laid out with absolute 
temperature and entropy for the coordinates. Steam at a certain 

condition of temperature and 
entropy is represented here 
by the point A. Then if 
some heat is added, increas- 
ing both temperature and 
entropy, the final condition 
is represented by the point 
B, and the area ABCD repre- 
sents the heat added in pass- 
ing from the condition at A 
to the condition at B. Such 
a diagram is usually called 
a temperature-entropy dia- 
gram, although the name 
heat diagram would prob- 
ably be more appropriate, since every area in expansions follow- 
ing a simple law represents a definite amount of heat. 

Another temperature-entropy diagram is shown in Fig. g } 
representing by the various shaded areas the heat added to water 
initially at the absolute zero of temperature to warm and com- 
pletely vaporize it at the temperature of evaporation of steam 
corresponding to the pressure Pi. The unshaded area under 
the irregular curve AA'B* represents the heat in a pound of 
water at the freezing point (3 2° F. or 492 in absolute tempera- 
ture). The area OBCD is the heat added to the water to bring 
it to the temperature of vaporization, or in other words, this 
last area represents the heat of the liquid (q) given in the steam 
tables for the pressure Pi. Further heating after vaporization 
begins is at the constant temperature Ti (corresponding to the 
pressure Pi) and is represented by an increasing area under line 
CE. When " steaming " is complete, the latent heat of evapo- 
ration is the area DCEF. If after all the water is vaporized 

* The area A "A 'BO represents the latent heat of fusion of ice, or it is the amount 
of heat that must be put into ice while melting, without producing a change of 
temperature. 



THE ELEMENTARY THEORY OF HEAT 



19 



more heat is added, the steam becomes superheated, and the 
additional heat required would be represented by an area to 




2.0 Entropy (0) 



Fig. 9. Temperature-Entropy Diagram showing the Total Heat in Dry Sat- 
urated Steam at the Temperature 7\ (measured from the Absolute Zero of 
Temperature) . 




.129 .566 1.528 Entropy (0) 

Fig. 10. Practical Example Illustrated with a Temperature-Entropy Diagram. 

the right of EF. (See Fig. 14.) The use of the temperature- 
entropy diagram in exhibiting the behavior of steam during 



20 THE STEAM TURBINE 

expansion will now be. discussed and illustrated with a practical 
example. 

Fig. 10 illustrates the heat process going on when feed water 
is received in the boilers of a power plant at ioo° F., is heated and 
converted into steam at a temperature of 400 F., and then loses 
heat in doing work. When the feed water first enters the boiler 
its temperature must be raised from ioo° to 400 F. before any 
" steaming " begins. The heat added to the liquid is the area 
MNCD. This area represents the difference between the heat 
of the liquid of steam at 400 F. (q c ) and at ioo° F. (q n ) and is 
about 306 B.T.U. The horizontal or entropy scale shows that 
the difference in entropy between water at ioo° and 400 F. is 
about .437* 

Every reader should understand how such a diagram is con- 
structed and especially how the curves are obtained. In this 
case the curve NC is made by plotting from the steam tables 
the values of the entropy of the liquid (usually marked with the 
symbol n or 6) for a number of different temperatures between 
ioo° and 400 F. and connecting the plotted points with a 
smooth curve. 

Now when water at 400 F. is converted into steam at that 
temperature, the curve representing the change is necessarily 
a constant temperature line and therefore a horizontal, CE. 
Provided the vaporization has been complete, the heat added in 
the " steaming " process is the latent heat of evaporation of 
steam (r) at 400 F., which, from the steam tables, is approxi- 
mately 827 B.T.U. 

The change in entropy during vaporization is, then, the heat 
units added (827) divided by the absolute temperature at which 
the change occurs (400 + 460 = 86o° F. absolute) or 

— = — - = .962 (see footnote, page 17). 
1 860 

The total entropy of steam completely vaporized at 400 F. is, 

* As actually determined from Marks and Davis' Steam Tables, pp. 9 and 15, 
the difference in entropy is .5663 — .1295 or .4368. Practically it is impossible to 
construct the scales in the figure very accurately. 



THE ELEMENTARY THEORY OF HEAT 21 

therefore, the sum of the entropy of the liquid (water) .566 and 
the entropy of the steam .962, or 1.528* To represent then 
this final condition of the steam, the point E is plotted on a 
horizontal line through C where the entropy measured on the 
horizontal scale is 1.528, as shown in the figure.f The area 
MNCEF represents then the total heat added to a pound of feed 
water at ioo° F. to produce steam at 400 F., and the area 
OBCEF represents, similarly, the total heat (H in the steam 
tables) in a pound of steam at 400 F. above that in water at 
32 F. 

Adiabatic Expansion and Available Energy. The practical 
example illustrated by Fig. 10 will also be used to explain how 
the temperature-entropy diagram can be used to show, for ideal 
conditions, how much work can be obtained by a theoretically 
perfect engine from the adiabatic expansion of a pound of steam. 
When steam expands adiabatically — without a gain or loss of 
heat — its temperature falls. Remembering that areas in the 
temperature-entropy diagram represent quantities of heat and 
that in this expansion there is no exchange of heat, it is obvious 
that the area under a curve of adiabatic expansion must be 
zero ; and this condition can be satisfied only by a vertical line 
which is a fine of constant entropy .J For the case in Fig. 10 
the expansion curve will He, therefore, along the vertical line 
EF, and if the temperature falls to ioo° F. the expansion will be 

* In all steam tables, entropy, like the total heat (H), the heat of the liquid (q), 
and the heat of evaporation (r), is measured above the condition of freezing water 
(32° F.). 

f The point E is shown located on another curve ST, which is determined by 
plotting a series of points calculated the same as E, but for different pressures. 
If more heat had been added than was required for vaporization, the area DCEF 
would have been larger and E would have fallen to the right of ST, indicating by 
its position that the steam had been superheated. The curve S T is therefore a 
" boundary line " between the saturated and superheated conditions. This curve 
can also be plotted from the values of temperature and entropy obtained from a 
table of the entropy of dry saturated steam. 

| Since in an adiabatic expansion there is no change of entropy, lines of constant 
entropy, in practice, are often called "adiabatics." It is very rare in steam turbine 
work that the expansion in a nozzle departs far from the adiabatic. For this reason 
other kinds of expansion are not mentioned here. 



22 



THE STEAM TURBINE 








c/ 


Ti and Pi ) 


,E 




^^^H 






ft 

1 


N^ 




To and P 2 


G \ 


£' 


B 





















F 


F' 



from E to G. During this change some of the steam has been 
condensed.* If now heat is removed from this mixture of steam 

and water as it exists at 
G till all the steam is re- 
duced to the liquid state, 
but without further lower- 
ing of the temperature, 
the horizontal line GN 
will represent the change 
in its condition. The 
quantity of heat absorbed 
in this last process — tech- 
nically known as condens- 
ing the steam — is rep- 
resented by the area 
MNGF, and the heat con- 
verted into work is, there- 
fore, the area NCEG ; and 
this is called the available energy. By means of diagrams like 
those in the preceding figures, it will now be shown how the 
available energy of dry saturated steam for any given conditions 
can be readily calculated from the data given in steam tables. 

Fig. ii is a temperature-entropy diagram representing dry 
saturated steam which is expanded adiabatically from an initial 
temperature Ti corresponding to a pressure Pi to a lower final 

* For the steam to be dry and saturated at the end of this process, the expansion 
would have been along the saturation line ET, and G would have appeared at G'. 

The heat of the liquid, q, of a pound of steam at ioo° F. is represented by OBNM, 
and the heat of evaporation (r) is MNG'F', so that the total heat (q + r or H) 
is OBNG'F'. The total heat of wet steam is expressed by q + xr, where x is the 
quality or relative dryness. In the case of this advabatic expansion, then, q is as 
before OBNM and xr is MNGF. It is obvious also that the lines NG and NG' 
have the same relation to each other as the areas under them, so that 



0=0 $1 02 Entropy 

Fig. ii. Temperature-Entropy Diagram for 
Steam Initially Dry and Saturated. 



line NG _ area MNGF _ xr 
line NG' ~ area MNG'F' ~ r 



NG 
° r NG' =X > 



showing that the quality of the steam at any point, G, on a constant temperature 
line (which for saturated steam is also a constant pressure line) is determined as 
in this case by the ratio of NG to NG'. 



THE ELEMENTARY THEORY OF HEAT 23 

temperature T 2 corresponding to a pressure P 2 . The other initial 
and final conditions of total heat H and entropy <£ (above 
3 2 F.) are represented by the same subscripts 1 and 2. The 
available energy or the work that can be done by a perfect en- 
gine under these conditions is the area NCEG. It is now de- 
sired to obtain a simple equation expressing this available energy 
E a in terms of total heat, absolute temperature, and entropy. 
Explanations of the preceding figures should make it clear that 

Hi = area OBNCEGF, 

H 2 = area OBNG'F', 

E a = area NCEG = areas (OBNCEGF + FGG'F') - OBNG'F', 

E a = Hx - H 2 + FGG'F', 

therefore E a = Hi - H 2 + (<fe - <£i) T 2 * (1) 

An application of this equation will be made at once to deter- 
mine the heat energy available from the adiabatic expansion of 
a pound of dry saturated steam at an initial pressure of 165 
pounds per square inch absolute to a final pressure of 15 pounds 
per square inch absolute. 

Example. Pi = 165 Ti = . . . 

P 2 = 15 T 2 = 673.0 from steam tables. f 

Hi = 1 195.0 from steam tables. 
H 2 = 1 1 50. 7 from steam tables. 
0i = 1. 56 1 5 from steam tables. 
<fa = 1.7549 from steam tables. 

Substituting these values in equation (1), we have 

E a = 1195.0 - 1150.7 + (i.7549 - i-5 6l 5) 6 73-° = I74-46 

B.T.U. per pound of steam. 
Now if in a suitable piece of apparatus like a steam turbine 
nozzle, all this energy that is theoretically available could be 

* It should be observed that this form is for the case where the steam is initially 
dry and saturated. For the case of superheated steam a slightly different form is 
required which is given on page 28. 

t The values of the properties of steam given in the exercises are taken from 
Marks and Davis' Steam Tables and Diagrams. 



24 



THE STEAM TURBINE 



changed into velocity, then we have by the well-known formula 

in mechanics,* for unit mass, 

V 2 

— = E (foot-pounds) = E a (B.T.U.) X 778, 

2g 



V = V778 X 2gE a = 223.7 Ve«, (2) 

where V is the velocity of the jet and g is the acceleration due to 
gravity (32.2), both in feet per second. 

Solving then for the theoretical velocity obtainable from the 
available energy in the practical example above, 

V = 223.7 V174.46 = 223.7 X 13.20 = 2953 feet per second.! 



1 

u 

p. 






c 


/ Ti and P t E'\e 








,} \ 




H 




n/ 




T2 and 


Pa 


gJg 


w 


33' > 


S 










1 

1 
1 
1 
1 
1 
1 
1 
1 

















1 

1 
1 

1 

»! 

FjF 


|F' 



= 



<f>x $1 $2 Entropy 



Fig. na. Temperature-Entropy Diagram for Wet Steam. 

The important condition assumed as the basis for the determi- 
nation of equation (1), that the steam is initially dry and satu- 
rated,* must not be overlooked in its application. There are, 
therefore, two other cases to be considered: 

(1) when the steam is initially wet, 

(2) when the steam is initially superheated. 

Available Energy of Wet Steam. Initially wet steam is easily 
treated in the same way as dry saturated steam. Fig. 11a is an 

* See Church's Mechanics of Engineering, p. 672, or Jameson's Applied Mechanics 
and Mechanical Engineering, vol. I, p. 47. 

t Losses in nozzles are neglected. A carefully made nozzle may have practically 
100 per cent efficiency. For discussion of nozzle losses see pages 57, 58, and 102. 



THE ELEMENTARY THEORY OF HEAT 25 

example of the case in hand. At the initial pressure Pi, the total 
heat of a pound of wet steam (qi + Xiri) is represented in this 
diagram by the area OBNCE"F". The initial quality of the 

CE" 

steam (xi) is represented by the ratio of the lines -7=-. The 

available energy from adiabatic expansion from the initial tem- 
perature Ti (corresponding to the pressure Pi) to the final tem- 
perature T 2 (corresponding to the pressure P 2 ) is the area 
NCE^G". If we call this available energy E aw , we have 

E aw = area OBNCEGF + FGG'F' - OBNG'F' - G"E"EG, 

E aw = Hi - H 2 + Ote - 0i) T 2 - (0! - 4> x ) (Ti - T 2 ) * 

E aw = Hi - H 2 + te - 0i) T 2 - £- (1 - xi) (Ti - T 2 ). (i0 

The velocity corresponding to this energy is found by substi- 
tution in equation (2), just as for the case when the steam was 
initially dry and saturated. 

Example. Calculations for the velocity resulting from adia- 
batic expansion for the same conditions given in the preceding 
example on page 23, except that the steam is initially 5 per 
cent, wet, are given below. 

Pi = 165 lbs. abs. Ti = 826.0 F. 
P 2 = 15 lbs. abs. T 2 = 673. o° F. 

Hi = 1195.0B.T.U. 

H 2 = 1150.7 B.T.U. 
0i = 1,5615. 

<fo = 1-7.549- 

ri = 856.8 B.T.U. 

xi = 1. 00 - .05 = .95. 

XT T\ 

* In general terms, = -= + 6. Here <£i = -=■ + 0i because x — 1. 
1 1 1 

<Px — -jT "T 01- 

01 — <t>X = jT (i — Xl). 



26 THE STEAM TURBINE 

Eav, = 1195-0 - 1150-7 + (i-7549 - i-5 6l 5) 673-0 - T~r~ 

020.0 

X .05 (826.0 - 673.0), 
E aw = 166.53 B.T.U. 
V = 223.7 VeZ = 223.7 X 12.90 = 2886 feet per second. 

It is observed that the theoretical velocity is reduced from 
2953 to 2886 feet per second by the presence of moisture in 
the steam. The percentage reduction in velocity is, however, 
only about 2 per cent, while the amount of moisture is 5 per 
cent. Moisture in steam produces still other and greater losses 
in turbines, which will be studied later. 

Available Energy of Superheated Steam.* In the following 
paragraphs the significance of the temperature-entropy curves 
for superheated steam will be explained, and it will be shown also 
how they are to be used to determine the available energy and 
the corresponding theoretical velocity resulting from adiabatic 
expansion in a nozzle. 

Specific Heat of Superheated Steam. In modern practice, 
superheated steam often enters our calculations and a trouble- 
some modification of the entropy diagram results. The difficulty 
arises because the specific heat of superheated steam is not at all 
accurately known. The diagrams in the appendix are calculated 
for the specific heat determinations by Knoblauch and Jakob. f 
The specific heat of steam varies with the temperature and pres- 
sure as shown in Figs. 12 and 13, giving values of the mean and 
the true specific heat at constant pressure (C p ). 

* In the following pages the important properties of superheated steam with 
which the modern engineer must deal will be briefly discussed. It is generally rec- 
ognized that a gain in steam economy results from the use of superheated steam in 
either steam turbines or reciprocating engines, but an accurate analysis of tests for 
the actual gain in economy of a plant is very difficult because there are so many 
factors entering. The peculiar circumstance, also, that water can exist indefinitely 
in the liquid state in the presence of superheated steam, makes conclusions from 
experimental data often uncertain. 

t Zeit. Verein deutscher Ingenieure, Jan. 5, 1907. Values of mean specific heat 
are taken from Mechanical Engineer, July, 1907, and Professor A. M. Greene's 
paper in Proc. American Society of Mechanical Engineers, May, 1907. 



THE ELEMENTARY THEORY OF HEAT 



27 



True specific heat represents the ratio of the amount of heat to 
be added to a given weight of steam at some particular condition 
of temperature and pressure to raise the temperature one degree 
to that required to raise the temperature of water at maximum 
density one degree. The mean specific heat is almost invariably 
used in steam turbine calculations. 









-|3 C 


_L_ 




1 


















4 U 


_t 






















3 AA- 


LI 






















I^-U- 


\a 
























t\\\ 


Xv 
























r v\ 


A\ 
























v X \ 


-Xt 
























H v v^ 


AXa 
























jA V 


N j£v^ 


















° d *R 








^\S3 


v 
















a. 9 * 5b 






/ \ s. 


V\S^ 


\ 






: n - 




« 






L X \ 


K^ 


\^s 


^J_ 






1 




M « 




4 


^v 


-sN^ 


^ s 










-284.4 

p245.8 

• 227.2 
-199.-0- 

•170.4 
-142.2^ 
L-113.6 

- 85.2 


Lbs. Abs 


3 <z> 

2 1 .54 




— / 


C— ^ 


s V. 


^c 








> -s 




.J 
















» 


a a 52 

e* s 
J M 
M -50 

18 




-A 




^— == 












» 


/ 


r 




"1 r 1 — pj 












-5G.8 

—28.4 
— 14.2 


» 


,46 


A- 












































































.42 



























200 250 300 350 400 450 500 550 600 650 700 750 
Temperature ° F 
Fig. 12. Mean Values of C p Calculated by Integration from Knoblauch and 

Jakob's Data. 

Temperature-Entropy Diagram of Superheated Steam. The 

graphic representation of the heat added during the superheat- 
ing of steam is easily accomplished with temperature-entropy 
diagrams. Fig. 14 shows a diagram similar to that representing 
dry saturated steam in Fig. 11, with the added area FEHJ to 
show the superheating from the temperature, Ti corresponding 
to the pressure Pi to the temperature of the superheated steam, 
T 8 . The total heat in a pound of steam above 32°F. is now 
represented by the area OBCEH s J. For adiabatic expansion of 
superheated steam at the temperature T 8 and pressure Pi to a 
pressure P 2 the available energy is the area LCEH 8 G. 



28 



THE STEAM TURBINE 



Too much calculation is involved in the construction of entropy 
diagrams to make a new diagram for every particular case from 
the properties usually found in steam tables; but the construc- 
tion of such diagrams should be understood.- From the expla- 



0.90 




100" 150 200 250 o 300 350 400 

Temperature C. 

Fig. 13. Values of the " True " Specific Heat of Superheated Steam. 

nations that have preceded, the construction of all the lines 
except EH S should be obvious. This line is obtained by calcu- 
lating the " change " or increase in the entropy of superheated 
steam for various values of temperature (between T a and Ti) from 
the following well-known relations: 



THE ELEMENTARY THEORY OF HEAT 



20 



CpdT 
T ' 



or 



Cf>s - 4>l = Cpm I l0g e — S I = 2.3028 Cpm ( logio T, - logio ^ J, 



where C^ is the mean value of the specific heat of steam taken 
from the curves in Fig. 12 for the temperature T 8 . 

The amount of energy that becomes available in the adiabatic 
expansion of superheated steam is very easily expressed with 




Fig. 14. Simple Temperature-Entropy Diagram for Superheated Steam. 

the help of Fig. 14a. Two conditions after expansion must be 
considered : 

(1) When the steam in the final condition is superheated, 

(2) When the steam in the final condition is wet (or dry 
saturated) . 

Using Fig. 14a with the notation as before except that E aa is 
the available energy from the adiabatic expansion of steam ini- 
tially superheated in B.T.U. per pound, 8 and H 8 are respec- 
tively the total entropy and the total heat of the superheated 



30 



THE STEAM TURBINE 



steam at the initial condition, then obviously from the diagram, 
when the steam is wet at the final condition, 



H s — H3 + (03 — S ) T3. 



(1") 



When the steam is superheated at the final condition, as, for 
example, when the temperature after expansion is T 2 ', 



E as — H s — H2 — (0 8 



02) T 2 .* 



(1"') 



It will be observed that these equations (1") and (i'") are the 
same in form as equation (1), page 23, and that equation (1") 
differs essentially in having the terms H s and 8 in the place of Hi 
and 0i. In other words equation (1) can be used for super- 
heated steam if the total heat and entropy are read from the 
steam tables for the required degrees of initial superheat. 

The following examples illustrate the simplicity of calculations 
with these equations : f 

Example. Steam at 150 pounds per square inch absolute 
pressure and 300 F. superheat is expanded adiabatically to 1 
pound per square inch absolute pressure. How much energy 
in B.T.U. per pound E as is made available for doing work? 
(Steam is wet in its final condition.) 

Solution. H s = 1348.8 B.T.U. per pound, 
H 3 = 1 103.6 B.T.U. per pound, 
03 = 1.980, 

0« = i-73 2 , 
T 3 = 559-6° F, 

E a8 = 1348.8 - 1103.6 - (1.980 - 1.732) 559.6 
= 383.9 B.T.U. per pound. 

The result above may be checked with the total heat-entropy 
chart in Marks and Davis' Steam Tables and Diagrams (Dia- 
gram I), and obtain (1349-967) or 382 B.T.U. per pound. 

* This is only approximately correct because the continuation of the line Tt to 
T a ' is not quite horizontal, and T s ' is usually unknown except as read from a chart. 

t Moyer and Calderwood's Engineering Thermodynamics, pages 125-128 (John 
Wiley & Sons, New York). 



THE ELEMENTARY THEORY OF HEAT 



31 



Example. Data same as in preceding example except that 
the final pressure is now 35 pounds per square inch absolute. 
(Final condition of steam is superheated.) Calculate E aa . 

Solution. H s = 1348.8 B.T.U. per pound, 
H 2 = 1 166.8 B.T.U. per pound, 
<j> s = 1.732, 
02 = 1.6868, 
T 2 = 718.9 F., 

E a8 = 1348.8 - 1166.8 - (1.7320 - 1.6868) 718.9 
= 149.5 B.T.U. per pound. 

Equations can also be written, as illustrated in Fig. 14, for 
the available energy E as for superheated steam when the final 
condition is " wet " thus, 

E a8 = Hi - (h 2 + x 2 r 2 ) + C pm (T. - Ti), 
or in a different form as, 

H\ — H 2 + C pm (T s — Ti) + (02 — 0s) T 2 . 
When at the final condition it is superheated, then 

E as = Hi — H 2 + Cpm (T 8 — Ti) — Cpm (T a f — T2), 

where other symbols are used as before and T 8 is initial tem- 
perature and T/ is the final temperature when superheated. 

The shaded area NCEG in Fig. 11 is also known as the the- 
oretical Rankine cycle * for the case where the steam supplied 
is initially dry saturated. The available energy E a , therefore, 
as given by equation (1) on page 23, multiplied by 778 gives the 
maximum theoretical foot-pounds of work that can be accom- 
plished with this cycle, neglecting losses, from a pound of dry 
steam. There are 33,000 X 60 foot-pounds in one horsepower- 
hour, and hence dividing 33,000 X 60 by E a X 778 f we get the 
theoretical steam consumption or theoretical " water rate " of 
a turbine using the ideal Rankine cycle with steam initially dry 
saturated. Similarly the area NCE^G" in Fig. 11a shows the 

* Moyer and Calderwood's Engineering Thermodynamics, pages 127-130. 
f The energy theoretically available for doing work in foot-pounds per pound of 
steam is E a X 778. 



32 



THE STEAM TURBINE 



available work for the theoretical Rankine cycle when the steam 
is initially wet, and the theoretical steam consumption of the 
Rankine cycle for this case is 33,000 X 60 divided by E aw * X 778. 

Fig. 14 shows also the Rankine cycle for steam initially super- 
heated. Calculation of theoretical steam consumption is similar 
to the cases already explained. 

The most important part of the design of a nozzle is the deter- 
mination of the areas of the various sections — especially the 




^1 ^2 0s ^s Entropy 

Fig. 14a. Temperature-Entropy Diagram for Superheated Steam. 

smallest section, if the nozzle is of an expanding or diverging 
type. Various forms of standard nozzles are shown in Fig. 15. 
In order to calculate the areas of nozzles we must know how to 
determine the quantity of steam (flow) per unit of time passing 
through a unit area. It is very essential that the nozzle is well 
rounded on the " entrance " side and that sharp edges along 
the path of the steam are avoided. Otherwise it is not impor- 

* From equation (1'), page 25. 



THE ELEMENTARY THEORY OF HEAT 



33 




\ ^mmm^ \ 




DeLaval Type. 




Nozzle 

Diaphragm 



Curtis Type. 

Fig. 15. Examples of Standard Designs of Nozzles. 

tant whether the shape of the section is circular, elliptical, or 
rectangular with rounded corners. Typical " square," " rectan- 
gular " and circular nozzle sections used in different makes of 
commercial turbines are shown in Fig. 15a. 



34 



THE STEAM TURBINE 



General Remarks Regarding Nozzles. Finally it may be 
stated that there is practically no difference in the efficiency of 
the nozzles used in commercial turbines if they have smooth 
surfaces and are properly designed for the correct ratio of the 








Fig. 15a. Sections of Nozzles Used in Commercial Turbines. 

area at the throat to that at the mouth, and if the length is not 
made much less than, nor more than possibly twice, that calcu- 
lated by the formula given on page 54. 

Whether the nozzle section is throughout circular, square, or 
rectangular (if these last sections have rounded corners) the 
efficiency as measured by the velocity will be about 96 to 97 per 
cent., corresponding to an equivalent energy efficiency * of 92 to 
94 per cent. Speaking commercially, therefore, it does not seem 
to be worth while to spend a great deal of time in the shops to 
make nozzles very exactly to some difficult shape. Simpler and 
more rapid methods of nozzle construction should be introduced. 
In some shops the time of one man for two days is required for 
the hand labor alone on a single nozzle. 

Example. Calculate the work done in foot-pounds by one 
pound of steam expanding behind a piston in a reciprocating 
engine for the conditions given in the example on page 23. 
(See discussion on page 16.) Ans. 174.2 X 778 ft.-lbs. 

Example. If the flow of dry steam at 165 pounds per square 
inch absolute pressure from a nozzle with a cross-sectional area 
of .00128 square foot is 2 pounds per second, what is the velocity 
of the discharging jet ? 

At the pressure stated steam has a specific volume of 2.75 

cubic feet per pound (from steam tables) . 

* Velocity is proportional to the square root of the available energy; and v .92 = 
.96, also V.94 = .97. 



THE ELEMENTARY THEORY OF HEAT 35 

Let V = velocity of discharge (ft. per sec.) 

A = area of nozzle = .00128 sq. ft. 
AV = volume discharged (cu. ft. per sec.) 
.00128 V = 2 X 2.75 

V = 4297 ft. per sec. 

If turbine blades could be made to transform all this velocity 
into useful work, how much horsepower could be transmitted to 
machinery from its shaft ? 

WV 2 

Kinetic energy of jet (ft.-lbs. of work per sec.) = > 

2g 

where W is the weight in pounds of the steam flowing and g 
is the acceleration due to gravity (32.2 ft. per sec). 

TT ft.-lbs. of work per sec. 
Horsepower = 

= 2 X (42Q7) 2 

2 X 32.2 x 550 
= about 1043 h.p. 

Example. What is the theoretical steam consumption (water 
rate) of the Rankine cycle for the conditions given in the ex- 
ample on page 23 ; that is, for steam initially dry saturated ? 

Am. 33,oooX6o . 

174.46 x 778 

Example. What is the theoretical steam consumption of the 
Rankine cycle for steam initially wet at the conditions stated 
in the example on page 25 ? 

. 3S,ooo X 60 

Ans. -"77— -. 

166.53 X 77s 



CHAPTER III. 

FLOW OF STEAM AND NOZZLE DESIGN. 

Flow of Dry Saturated Steam through Nozzles. The weight 
of steam discharged through any well-designed nozzle with a 
rounded inlet, similar to those shown in Figs. 13 and 14, depends 
on the initial absolute pressure (Pi), if the pressure against which 
the nozzle discharges (P 2 ) does not exceed .58 of the initial pres- 
sure. This important statement is well illustrated by the follow- 
ing example. If steam at an initial pressure (PJ of 100 pounds 
per square inch absolute is discharged from a nozzle, the weight 
of steam flowing in a given time is practically the same for all values 
of the pressure against which the steam is discharged (P 2 ) which 
are equal to or less than 58 pounds per square inch absolute. 

If, however, the final pressure is more than .58 of the initial, 
the weight of steam discharged will be less, nearly in proportion 
as the difference between the initial and final pressures is reduced. 
(See pages 41 and 42.) 

The most satisfactory and accurate formula for the " constant 
flow" condition, meaning when the final pressure is .58 of the 
initial pressure or less, is. the following, due to Grashof,* where F 
is the flow of steam f (initially dry saturated) in pounds per 

* Grashof, Theoretische Maschinenlehre, vol. i, iii; Hiitte Taschenbuch, vol. i, 
page $$$. Grashof states the formula, 



F = .01654 A Q P r 



• 9696 



but the formula given in equation (3) is accurate enough for all practical uses. 

f Napier's formula is very commonly used by engineers and is accurate enough 
for most calculations. It is usually stated in the form 

F = A -^> 
70 

where F, P t , and A Q have the same significance as in Grashof s formula. The 
following formula is given by Rateau, who has done some very good theoretical 

36 



NOZZLE DESIGN 



37 



second, A is the area of the smallest section of the nozzle in square 
inches, and P t is the initial absolute pressure of the steam in 
pounds per square inch, 

A p .97 

or, in terms of the area, 

Ao=||- (3') 

These formulas are for the flow of steam initially dry and 
saturated. An illustration of their applications is given by the 
following practical example. 

Example. The area of the smallest section (A ) of a suitably 
designed nozzle is .54 square inch. What is the weight of the 
flow (F) of dry saturated steam per second from this nozzle when 
the initial pressure (P x ) is 135 pounds per square inch absolute 
and the discharge pressure (P 2 ) is 15 pounds per square inch 
absolute? 

Here P 2 is less than .58 P x and Grashof's formula is applicable, 
or, 

F = 



54 (i35)- 97 



60 

„ .54 X 116.5* , , 

F = — — - = 1.040 pounds per second. 

60 

Flow of Wet Steam. When steam passes through a series of 
nozzles one after the other as is the case in many types of tur- 
bines, the pressure is reduced and the steam is condensed in 
each nozzle so that it becomes wetter and wetter each time. In 
the low-pressure nozzles of a turbine, therefore, the steam may 

and practical work on steam turbines, but his formula is too complicated for 
convienent use: 

F = .001 A P l [15.26 - .96 (log P x + log. 0703)]. 

Common or base 10 logarithms are to be used in this formula. 

60 
* A curve from which values of 5-^ can be read is given on page 47 (Fig. 19). 

*Y 

The flow (F) calculated by Napier's formula for this example is F — — — , 

or 1. 041 pounds per second. 



38 THE STEAM TURBINE 

be very wet although initially it was dry. Turbines are also 
sometimes designed to operate with steam which is initially wet, 
and this is usually the case when low-pressure steam turbines 
(see Chapter IX) are operated with the exhaust from non-con- 
densing reciprocating engines — a practice which is daily be- 
coming more common. In all these cases the nozzle area must 
be corrected for the wetness of the steam. For a given nozzle 
the weight discharged is, of course, greater for wet steam than 
for dry; but the percentage increase in the discharge is not 
nearly in proportion to the percentage of moisture as is often 
stated. The general equation for the theoretic discharge (F) 
from a nozzle is in the form * 



: flow is 



F =K 

* The general equation for the theoretic flow is 



is 

the pressure at any section of the nozzle, v x is the volume of a pound of steam at 
the pressure P v and k is a constant. The flow, F, has its maximum value when 

2 fc+i 

kfc fp\ k 



sr-ea 



is a maximum. Differentiating and equating the first differential to zero gives 

Px U + 1/ 
P 2 is now the pressure at the smallest section, and writing for clearness P for 



Now regardless of what the final pressure may be, the pressure (P ) at the smallest 
section of a nozzle (A ) is always nearly .58 P x for dry saturated steam. Making 
then in the last equation P = .58 P x and putting for k Zeuner's value of 1.135 
for dry saturated steam, we may write in general terms the form stated above, 

where K is another constant. See Peabody's Thermodynamics of the Steam 
Engine, page 132; Zeuner's Theorie der Turbinen, page 268 (Ed. of 1899). 



NOZZLE DESIGN 39 

where P x is the initial absolute pressure and Vi is the specific 
volume (cubic feet in a pound of steam at the pressure Pi). 
Now, neglecting the volume of the water in wet steam, which is 
a usual approximation, the volume of a pound of steam is pro- 
portional to the quality (x\). For wet steam the equation above 
becomes then 

F = K Jfl . 

The equation shows, therefore, that the flow of wet steam is 
inversely proportional to the square root of the quality (xi). 
Grashof's equations can be stated then more generally as 



(4) 



. oor vxi , ,x 



60 Vxi 

60F Vx^ 
Pi 



These equations become, of course, the same as (3) and (3') 
for the case where X\ — 1. 

Flow of Superheated Steam through Nozzles. The discharge 
of superheated steam from a nozzle is one of the most important 
subjects of which the engineering profession generally has no 
correct data. The author has observed in his practice again and 
again that the formulas ordinarily given for the flow of super- 
heated steam were not correct and more reliable data had to be 
found. The formulas given here were actually determined from 
the data of Lewicki's experiments with a 30-horsepower De Laval 
turbine * but were later checked with a great mass of data in the 
possession of the General Electric Company. The precision 
with which the formula applies to Lewicki's data is shown in the 
table given on the next page. 

A formula was desired to express the flow of superheated steam 
discharged from a nozzle in the form of formula (3) for the flow 
of dry saturated steam, together with a suitable coefficient to 

* Zeit. Verein deidscher Ingenieure, April 4, 1903, page 494. 
Mitteilungen tiber Forschungsarbeiten, Heft 12 (1904), Zalentafel 25. 



40 



THE STEAM TURBINE 



correct for the effect of superheat. A formula of this form is 
expressed by 



F = 



60 (1 + .00065 D) 
60 F ( 1 + .00065 D) 

Pi -97 



j or 



(5) 
(SO 



where F is the weight in pounds of superheated steam discharged 
per second, A is the area of the smallest section of the nozzle in 
square inches, Pj is the initial pressure in pounds per square inch 
absolute, and D is the superheat in degrees Fahrenheit.* 

Lewicki's data for the tests f given below were in metric units 
but are recorded here in the corresponding English units. 

Initial pressure P t = 99.25 pounds per square inch absolute. 

Final pressure P 2 = 14.6 pounds per square inch absolute. 



Number of Test. 



Temperature of steam, degrees F 

Superheat, degrees F. (D) 

Flow, pounds per hour (tests) . . 

(1+ .00065 D) 

Flow, pounds per hour, corrected 
by formula (5) to equivalent 
flow of dry saturated steam . . 



I 

386.6 

59-6 

882.0 
1.038 


2 

463.1 
136. 1 

837.0 
1 .089 


3 

491.9 
164.9 
824.0 
1. 107 


4 


5 


6 

619.7 
292.7 
776.O 
I. 190 


7 


529-7 
202. 7 
804.0 
1-132 


592.7 
265.7 

778.0 
1. 173 


7°3-4 
37 6 -4 
735-0 
1.245 


917.0 


910.0 


911. 


910.0 


914.0 


910.O 


914.0 



723.6 
396.6 
729.0 
1.258 



916.0 



Volume of Superheated Steam. Thermodynamic relations 
show that the flow of superheated steam is inversely proportional 
to the square root of the specific volume, f so that from the 
author's equation for the flow of superheated steam (5) the fol- 
lowing formula for the specific volume is easily obtained : 

v s = (i + . 00065 D) 2 v, (6) 

* It is stated that Mr. A. R. Dodge has shown practically the same results from 
the Newport tests of a Curtis turbine reported by Mr. G. H. Barrus. 

t Mitleilungen iiber Forschungsarbeiten, Lewicki, Heft 12 (1904), Zalentafel 25. 

t This relation is discussed by the author in Mechanical Engineer (London), 
Aug. 24, 1907, page 277, and in the Harvard Engineering Journal, June, 1907, 
page 36. Compare with Stodola, Die Dampfturbinen, 3rded., page 9. 

This formula gives values of specific volume representing a fair average of results 
obtained from the formulas of Zeuner, Tumlirz, Knoblauch, and Schmidt (based 
upon Hirn's experiments). 



NOZZLE DESIGN 



41 



where v a is the volume in cubic feet of one pound of superheated 
steam, v is the volume in cubic feet of one pound of dry saturated 
steam at the same pressure as for v., and D is the superheat in 
Fahrenheit degrees. 

Flow of Steam when the Final Pressure is more than .58 of the 
Initial Pressure. For this case the discharge depends upon the 
final pressure as well as upon the initial. No satisfactory formula 
can be given in simple terms, and the flow is most easily calculated 
with the aid of the curve in Fig. 16 due to Rateau. This curve 



1.0 










































\ 




















































" aT 












.9 










































S 




















































Q. 












.8 




















y 


y 
























































































"* "S 7 

C 3 .1 

gg 6 




































































































































































§r 
























































•o S .5 
2 2? 














































































































Coefficient 
Formulas ( 

co rf^ 




































































































































































/ 






















































.2 


/ 






















































/ 






















































.1 








































































































































































1.0 



.8 .7 .6 

Ratio of Final to Initial PressureJjL 
Pi 



Fig. 16. Coefficients of the Discharge of Steam when the Final Pressure is 
Greater than .58 of the Initial Pressure. 



is used by determining first the ratio of the final to the initial 

p 
pressure — , and reading from the curve the corresponding co- 

Pi 
efficient showing the ratio of the required discharge to that cal- 
culated for the given conditions by either of the equations (3) 
or (4). The coefficient from the curve times the flow calculated 
from equations (3) or (4) is the required result. Obviously the 



42 THE STEAM TURBINE 

discharge for this condition is always less than the discharge 
when the final pressure is equal to or less than .58 of the initial. 
The actual design of the nozzles for a commercial turbine will 
be taken up in the next paragraph; but before this is done, one 
other equation used almost continually in nozzle and blade de- 
signs must be explained. It is to find the quality of the steam 
after an adiabatic expansion. The initial quality of the steam is 
usually determined by the conditions in the boiler equipment, or 
is given in the engineer's specifications for a new design, but the 
quality of the steam after each expansion must be calculated. 
The general equation for adiabatic flow (constant entropy *) is 

1 1 ±2 

and solving, 

*-[f +*-*]$' ' w 

where the subscript 1 attached to the symbols refers to the initial 
condition, and the subscript 2 to the final. The terms 0i and 2 
are the entropies of- the liquid (water) at the initial and final con- 
ditions, and the other symbols are used as before. 

To avoid the laborious calculation of equation (7) to deter- 
mine the quality after adiabatic expansion, curves of steam 
quality have been calculated and plotted on the entropy-total 
heat chart in the appendix. To illustrate the use of these curves 
an example is given below. 

Example. Steam at 165 pounds per square inch absolute 
pressure (Pi), which is 4 per cent, wet (x± = .96), is expanded 
adiabatically in a nozzle to 15 pounds per square inch absolute 
(P 2 ) • What is the quality after expansion ? 

Method. A point is first located on the chart where the quality 
curve for x = .96 crosses the pressure line for 165 pounds as shown 
diagrammatically in Fig. 17. A horizontal fine of constant 
entropy drawn through this point shows at its intersection with 
the pressure line for 15 pounds the quality after expansion. In 

* See footnote on page 17. 



NOZZLE DESIGN 



43 




Fig. 17. Illustrates the Use of the 
Entropy-Total Heat Chart to Deter- 
mine the Quality of Steam after Ex- 
pansion. 



this case the quality is .837. For practical designing to get 
satisfactory results the quality 
should be read to three signifi- 
cant figures. 

Nozzle Calculations. In the 
calculations to determine the 
dimensions of a nozzle it is 
necessary to have given the 
following data: 

(1) the weight of steam that 
is to be delivered through the 
nozzle to develop the required 
power in the turbine. 

(2) the initial and final pressure (Pi and P 2 ). 

(3) the quality (xi) of the steam supplied. 

With these data A Q is then calculated by substitution of these 
quantities in equation (4'). This is the area at the smallest 

section or throat as shown 
in Fig. 18. 

The area of the nozzle can 
be determined by simple cal- 
culations only at the small- 
est section or throat. To 
determine the area at any 
other section of the ex- 
panding portion between the 
throat and the mouth in- 



Throat 




ifouth 



TTTJ77 



Fig. 18. 



A Typical Expanding Nozzle 

volves equations of the form of those at the bottom of page 38. 
It is therefore convenient to determine the sections other than 
the throat by a proportional method. Now the areas of different 
sections depend on the following three conditions: 

(1) the velocity of the steam. 

(2) the specific volume. 

(3) the quality or dryness. 

The essential condition to observe is that the weight of steam 



44 THE STEAM TURBINE 

flowing per second is the same at every section; and for the same 
flow the areas are inversely proportional to the velocities at any 
two sections compared, and directly proportional to both the 
specific volumes and the qualities. We may then write the 
equation 

J»-J»X*X*, (8) 

A V x v x 

A = area of nozzle at the smallest section in square inches. 

A x = area of nozzle at any section of expanding portion in 
square inches. 

V = velocity of steam at the smallest section in feet per sec- 
ond* 

V x = velocity of steam at any section in feet per second* 

v = specific volume at the smallest section in cubic feet per 
pound. 

v x = specific volume at any section in cubic feet per pound. 

x = quality of steam at the smallest section. 

x x = quality of steam at any section. 

The product — - X - a X — when calculated for the largest 
V x v x 

section or mouth is often called the expansion ratio (see Fig. 21, 

page 49), and is very nearly proportional to the ratio of the 

initial to the final pressure. 

An example will now be given to show how the actual area of 
the nozzles of a commercial turbine can be calculated. 

Example. A test of a De Laval turbine was as follows: 

Pressure in the steam-chest (Pi) 21 1.5 pounds absolute 
Vacuum referred to 30-in. barometer. . 26.6 in. mercury 

Moisture in steam. 2.2 per cent. 

Brake horsepower 333 

Steam consumption, per brake horsepower- 
hour as weighed (" wet ") I 5-5 I pounds 

Number of nozzles open 8 

* Since practically all the loss in a nozzle occurs before the steam " emerges " 
from the throat, the same coefficient applies to both V and V x and cancels when 
expressed in equation (8). The non-expanding nozzles shown on page 59 are no 
more efficient than equally well made expanding nozzles. 



NOZZLE DESIGN 45 

In this case P 2 is given as 26.6 inches vacuum, which is less 
than 2 pounds absolute pressure, and is therefore less than .58 Pi 
and formula (4') is applicable, so that the " throat " area of the 
eight nozzles is expressed by 

a _ 6° F Vxi 
Ao ~ iY 97 ' 

where Xi = .978, Pi = 21 1.5 pounds per square inch absolute, 
and 

F = 233__ — >i_ = 51-1 p 0unc i s we t steam per second. 
3600 3600 

(211. 5)- 97 3600 

A = -333 X ^— ~ X .989 = .472 square inches. 
3600 

The area of the throat of each nozzle is therefore .0590 square 
inches. 

The value of — was read from the curve * of -— — in 

(2II.5)- 97 Pi- 97 

Fig. 19. 

The nozzles of most commercial types of steam turbines are 
made with straight sides as shown in Fig. 18, so that in addition 
to the area at the throat only one other area must be found to 
fully determine the expanding portion. This is obviously most 
easily determined at the mouth; since the velocity must be calcu- 
lated from the available energy for an adiabatic expansion from 
Pi = 2 1 1. 5 pounds per square inch absolute to P 2 = 1.67 pounds 
per square inch absolute (26.6 inches vacuum). This available 
energy can be calculated by equation (Y) for initially wet steam, 
but the calculation is laborious, and instead the energy will now 
be read from the entropy-total heat chart in the appendix. The 
point is first located on the chart where the line for 21 1.5 pounds 
pressure crosses the .978 steam quality line (estimated). Read- 
ing the scale of abscissas at this point we find that the total heat 
energy in a pound of steam at this condition is 1181 B.T.U. By 

* The curve was made in this form to make the final form of the result more 
convenient for slide-rule or cancellation calculations. 



46 THE STEAM TURBINE 

following a horizontal line from this point across the chart as 
indicated diagrammatically in Fig. 20 till it intersects the pres- 
sure line corresponding to 1.67 pounds (estimated), the total heat 
energy escaping with the exhaust steam after adiabatic expansion 
as read on the scale of abscissas is 874 B.T.U. The difference 
between the two readings, or 307 B.T.U. , is the available energy 
(E aw ). The quality at the end of expansion (x 2 ) as read from the 
curves is .767. In this way the labor of calculating x 2 is saved. 
From the value of the available energy due to expansion, E aw , 
the velocity V 2 at the mouth of the nozzle is calculated by equa- 
tion (2), or 

V 2 = 223.7 VEaw = 223.7 V307 = 3919 feet per second. 

In order to determine the ratio of the area at the mouth of the 
nozzle (A 2 ) to that at the smallest section (A ) by equation (8) 
the velocity (V ) and the quality (x ) * must be determined. 
These evaluations are most easily made in the same way as for 
V 2 and x 2 by means of the entropy-total heat chart. Now the 
available energy E a0 , corresponding to the velocity V Q , must be 
calculated for adiabatic expansion from Pi = 21 1.5 pounds and 
Xi = .978 to P = .58 Pi f = 122.7 pounds. This available energy 
is 44 B.T.U. and x is .939. The velocity V is, therefore, 
223.7 ^E a o = 223.7 V44 =1483 feet per second.^ 

* For steam initially dry and saturated, the quality after adiabatic expansion 

(x 2 ) for all practical cases is very nearly expressed, empirically, by the equation 



X-2 



-m 



and the quality at the throat (x ) may be taken as .965 for all practical cases re- 
gardless of the initial and final p essures. 

t It is well established by thermodynamic calculations and by actual experiment 
that the pressure P at the smallest section of a nozzle is always very nearly .58 of 
the initial pressure (Pi) . 

t Very elaborate curves of the velocities resulting from the adiabatic expansion 
of dry saturated steam have been prepared and published in some American books. 
Considering the several stages in nearly all types of turbines, such curves can be of 
very little use to practical men, because the condition that the steam admitted to 
the nozzles is dry and saturated occurs infrequently. That some of the authors 
neglected to mark the curves " for steam initially dry and saturated " deserves 
severe criticism. The curves, as given, are very misleading, as they are apparently 
intended for general application for all qualities and superheats. 



NOZZLE DESIGN 



47 






Initial Pressure, P a (5 to 50 Lbs.) 
5 10 15 20 25 30 3-5 40 45 



L2 



1.1 



^ 1.0 















EE 
















EE 






— ; 


= 




Tj|t 


H 






EH 


EE 




n 


= 


EE 


he 


= 


EEE 




_ 




ll 


|| 






= 


n 


E=E 








~ 




1 


1 


8 


n 


JEE 










^P 




















































H 



60 80 ,100 120 140 160 180 200 220 2i0 
.Initial. Pressure, P L (50 to 260 L,bs.) 



Fig. 19. Curves Showing Values of 



60 

p.V 




Total Heat (H) 



4 B.T.U. 



Fig. 20. Illustrates the Use of the Entropy-Total Heat Chart for Deter- 
mining the Available Energy in a Pound of Steam. 



48 THE STEAM TURBINE 

In addition to the values already obtained it is only necessary 
to get v and vi (the specific volumes of dry saturated steam at the 
corresponding pressures P and Pi) to determine all the terms 
in the equation for the expansion ratio as already given, and 
putting now the subscript 2 for x in equation (8) to express the 
conditions corresponding to the pressure P 2 ; then 

f-l&xSxS, (so 

A V 2 v x 
or 

. 148s w 206.0 vv .767 w • 1 / 

A 2 = - JX - i * X — : — X - J — L X .0590 = 1.030 square inches (area 
3919 3.642 .939 

at mouth) . 

The author has found as the result of some investigations 

regarding the design of nozzles that the expansion ratio (— ^ J of a 

properly designed nozzle is very nearly proportional to the ratio of 
the initial pressure (Pi) to the final pressure (P 2 ). The curve 
shown in Fig. 21 has been calculated on this basis for widely 
different conditions but for rather small expansions, and has been 
found to be accurate enough for practical purposes in designing 
turbines of more than one stage. A similar curve is now being 
used by .the nozzle designers of one of the large manufacturing 
companies. After the relations shown by Fig. 21 had been 
worked out, it was found that Zeuner had arrived at a similar 
result mathematically after making certain assumptions; * but 

* In Zeuner's Theorie der Turbinen, page 270, the following equation is given 
to express the ratio of the area at the mouth to that at the smallest section (expan- 
sion ratio) : 

M -1550 



A ° vgr-o 1 - 



where the terms A and P are used as in the equations above. There is probably 

)tic 

P-2 



some error in Zeuner's assumptions, because actually values of - are not quite 

#0 



constant for varying values of _ 



NOZZLE DESIGN 



49 



Zeuner's equation itself is practically useless on account of being 
too complicated.* 

Shapes of Expanding Nozzles. The inside walls of the expand- 
ing portion of the nozzle are usually surfaces with straight-line 































































28 




























































— 




























































26 


























































































































24 


























































































































22 


























































































































20 

s 


















































































































































































































































lie 

g 


























































































































■314 
a 


























































































































ft 
ol2 






















/ 




















































































gio 






































































































o 
2 8 
































































































For Steam Initially- 
Dry Saturated or Wet 










*G 










































































































4 
















/ 


























































/ 
















































2 












/ 
























































J— 





P 


= .5> Pi 








































Q 























































12 3 4 5 

Ao 
Expansion of Nozzle ,—55" 

Fig. 21. Curve Showing the Approximate Relation between Expansion Ratio 
of a Nozzle and the Ratio of the Initial to the Final Pressure. 

elements, meaning that in any section of the nozzle along the 
axis, like Fig. 18, the inner walls are shown by straight lines. 

* The author's curve in Fig. 21 is expressed by, 

t = -(£>- 

A more accurate form for pressure ratios greater than 25 is the following: 

t-n-ffiT- 



5° 



THE STEAM TURBINE 



Parenty has shown that, for the highest efficiency, theoret- 
ically, such a section along the axis should be slightly elliptical 
wiih the focus in the throat, but practically this shape shows no 
advantage and is much too difficult to construct. For making 
nozzles like those in Curtis turbines, where the work is done 
largely with hand tools, the construction of even the simplest 
form is very expensive and the cost of an elliptical curvature is 
practically prohibitive. The shape to give the best expansion 
curve has been the subject of investigation by various experi- 
menters.* As the practical results are particularly interesting, 
it may be well to describe briefly a typical form of apparatus 
usually employed in these experiments as shown in Fig. 22. The 




Fig. 22. Searching Tube Apparatus for Determining the Pressures in Nozzles. 



nozzle to be tested is marked A in the figure. The steam entering 
the passage B discharges through the nozzle directly into the 
exhaust pipe E. A small "searching" tube, C, is provided 
which is sealed at one end and has a very small hole, D, a short 
distance from this end. The other end of the tube is attached to 
a mercury column or pressure gauge. Suitable means are provided 
for sliding the " searching " tube with its pressure gauge back and 

* The conditions of pressure and velocity of steam inside a nozzle are discussed 
very completely from the mathematician's viewpoint in Die Dampfturbinen by 
Stodola, 3rd edition, pages 42 to 75, and in Zeitschrift fur das Gesamte Turbinen- 
wesen, Aug. 10, 1906, pages 325-327. 



NOZZLE DESIGN 



51 



forth so that pressures can be observed in different parts of the 
nozzle, corresponding to the position of the hole D. From these 
observations a curve of 
pressures may be made, 
and from this, together 
with the data of the weight 
of steam passing per unit 
of time, a second curve may 
be developed showing the 
corresponding velocities. 
The curves in Figs. 23, 24, 
and 25 are examples of the 
results obtained by this 
method for three very dif- 
ferent nozzles. The nozzle 
shown at the top of Fig. 
23 has curved lines, nearly 
elliptical, for its inside 
walls. A pressure curve 
is shown beneath the sec- 
tion drawing, in its true 
relative position corre- 
sponding to points along 
the axis of the nozzle. 
Theoretically this shape of nozzle approaches the ideal for an 
adiabatic expansion.* Practical conditions, however, as stated 

* A nozzle with a circular section (perpendicular to the axis) has hss surface 
exposed to the flow of steam than a nozzle of any other form of the same length 
and expansion. For this reason this form should give minimum friction losses. 
In practice, however, this type is not often used when the section at the mouth of 
the nozzle is made rectangular, at least when the nozzles are arranged in groups 
with the mouths of the several nozzles close together. There are obvious advan- 
tages from this last construction, as first pointed out by Professor Riedler, because 
if the nozzle mouths are made rectangular and close together a long continuous 
"band" of steam is secured which is approximately homogeneous and of constant 
velocity. The flow from the end nozzles is, of course, affected by excessive eddying 
and other irregularities just as single nozzles. Efficiency of the end nozzles is 
therefore considerably less than that of any of the others in the group. 




Fig. 



Distance along Axis of Nozzle 



23. Expansion Curve of a Nozzle 
with an Elliptical Axial Section. 



52 



THE STEAM TURBINE 



before, make the nozzle shown in Fig. 24 with expanding straight- 
line walls preferable if the throat and mouth areas are properly 
designed. Fig. 24, however, is intended to show primarily the 
effect of using a nozzle for non-condensing service, which was 
designed to be used condensing. For this reason the expansion 

in the nozzle is greater 
than it should be for the 
pressures with which it 
is operating; and for this 
reason th^ pressure inside 
the nozzle, as illustrated by 
the curve, falls below the 
exhaust pressure. This 
is called over-expansion 
or " over-compounding " 
and is always accompa- 
nied by a loss in efficiency. 
In fact, as will be shown 
again later, the effect of 
over-expansion, or making 
a nozzle too large at the 
mouth, reduces nozzle effi- 
ciency much more than if 
it is made the same per- 
centage too small. (See 
Fig. 28.) The curves 

> Distance along Axis of ^Nozzle in Fig. 24 show that the 

Fig. 24. Expansion Curve of a Nozzle with pressure at the mouth 
Straight Walls. . g ^ j.^ bwer than the 

atmospheric exhaust, and a partial vacuum is thus secured 
at the blades opposite the nozzles. When such nozzles are 
operated non-condensing there is some gain from the reduc- 
tion of disk and blade friction because the wheel and blades 
revolve in a less dense medium; but when considering also 
the increased losses in the nozzle itself because of over- 
expansion, there is certainly no net gain over having a nozzle 











140^ 
130 

£ 120 










£100 

% 

M 80 
u 

a> 
P* 

m 70 

to 
rt 
g 60 

S 50 

S 40 

30 

20 




















\ Atmospheric 1 Line 1 


10 



^^ / ~~ 


^^ 



NOZZLE DESIGN 



53 



designed exactly for the expansion corresponding to the operat- 
ing conditions. 

Fig. 25 is intended to show an abnormal but interesting form of 
nozzle which gives some idea of the behavior of steam when the 
expansion is not gradual and continuous. It was argued by a 
designer who made this nozzle that this form should be as efficient 
as any other. It was 



his theory that if the 
areas at the throat 
and at the mouth 
were of the right size, 
the shape of the walls 
between was of no 
consequence, and, in 
fact, that the steam 
of itself would take 
the correct passage. 
Thus by preventing 
the steam particles 
from touching the 
walls the friction 
losses in the nozzle 
should be reduced. It 
will be observed, how- 
ever, from the curve 
in the figure deter- 
mined from some experiments with this nozzle that the pres- 
sure first drops abruptly in the throat to .58 of the initial, 
as in any other nozzle, and then forms a series of waves, from 
which it appears that the particles of steam strike the walls and 
rebound, to meet again at a point, as at A, where an increased 
pressure is produced, and so on till the mouth is reached. ' The 
probable path of the steam is shown by the dotted lines in the 
drawing of the nozzle. These experiments show therefore that 
the steam will not take the correct passage through a nozzle with- 
out the provision of properly designed walls of gradually increasing 




Fig. 25. 



— >- Distance along Axis of NozzJe 
Expansion Curve in an Abnormal Nozzle. 



54 THE STEAM TURBINE 

area corresponding to the expansion required. The importance 
of careful workmanship in the manufacture of nozzles is therefore 
obvious. 

The results shown by Fig. 25 bring up naturally the discussion 
of the proper length for a nozzle, as the one in this figure was 
obviously much too long. 

Probably the best designers of the Curtis types of turbines make 
the length of the nozzle depend only on the initial pressure. In 
other words, the length of a nozzle for 150 pounds per square inch 
initial pressure is usually made the same for a given type regard- 
less of the final pressure. And if it happens that there is 
crowding for space, one or more of the nozzles is sometimes 
made a little shorter than the others. 

Designers of De Laval nozzles follow practically the same 
"elastic" method. The divergence of the walls of non-con- 
densing nozzles is about 3 degrees from the axis of the nozzle, and 
condensing nozzles for high vacuums may have a divergence of 
as much as 6 degrees * for the normal rated pressures of the 
turbine. 

The author has used successfully the following empirical 
formula to determine a suitable length, L, of the nozzle between 
the throat and the mouth (in inches): 

L = V15 A , (9) 

where A is the area at the throat in square inches. 

The design of the nozzle calculated in the example on page 36 
can now be completed with the determination of its proper length, 



L = V15 X .059 = .9 inch. 

The important dimensions of nozzles of circular section suitable 

* According to Dr. O. Recke, if the total divergence of a nozzle is more than 
6 degrees, eddies will begin to form in the jet. There is no doubt that a too rapid 
divergence produces a velocity loss. 

When a number of nozzles intended for different initial pressures are supplied 
for use in the same turbine, the length as determined by the taper is usually made 
to correspond to the pressure that is to be most used. Inspection of the De Laval 
nozzle in Fig. 15 shows that it is necessary to make all the nozzles of the same length 
for a given size of De Laval turbine, so that the nozzles maybe used interchangeably. 



NOZZLE DESIGN 55 

for this De Laval turbine tested by Dean & Main may be tabu- 
lated as follows: 

Area at throat (A ), .0590 square inch. Diameter (O ), .274 
inch. 

Area at mouth (A 2 ), 1.008 square inches. Diameter (D 2 ), 
1. 132 inches. 

Length of nozzle (L) as determined by equation (9), .9 inch. 

Length of nozzle assuming a divergence of 12 degrees, 1.9 
inches. 

It will be observed from the last calculation that a de- 
signer of De Laval nozzles would make the length about twice 
that calculated by equation (9). The nozzles of De Laval 
turbines are made unusually long largely for mechanical rea- 
sons. There is probably very little loss in this additional 
length. 

A nozzle of circular section suitable for these conditions is 
shown at the top of page 33 (Fig. 15). It will be observed that 
a rounded entrance to the nozzle has been made, If a well- 
rounded entrance is not provided the rate of flow through the 
nozzle may be only 50 to 70 per cent, (depending of course on the 
sharpness of the corners) of the normal flow calculated from 
Grashof's formulas given in equations (3) and (4). The effi- 
ciency is also very much reduced if the steam is not led to the 
throat along a surface of gradual curvature.* 

* Jude states that a very large rounded inlet appears to "choke" the nozzle 
a little. He admits that it gives maximum discharge "but at the expense of kinetic 
energy, that is, of the kinetic energy effective in an axial direction." The results 
of Rateau's experiments seem to show, however, that the efficiency of a convergent 
nozzle suitably rounded is unity. If any loss does result from a rounded entrance 
which is too large it is probably of negligible amount. Some conclusions drawn 
from Rosenhain's experiments reported in Proc. Inst. Civil Engineers, vol. 140, 
may be of interest in this connection. A series of experiments was made with vari- 
ous nozzles working from 20 pounds to 200 pounds per square inch gauge pressure 
with atmospheric exhaust. The most efficient form of nozzle up to about 80 pounds 
gauge pressure appears to be a plain orifice in a thin plate, as measured by nozzle 
reaction (see page 60), but this does not imply that such a form is the best nozzle 
for a turbine under similar conditions. With this kind of orifice there is too much 
spreading of the jet, and the internal eddies and whirls are too violent for useful 
application at the point where the jet strikes the turbine vanes. 



56 



THE STEAM TURBINE 




It has been shown by Stodola's experiments that the difference 
in pressure between the outer and inner portions of the jet inside 
a nozzle of approximately correct design are practically negligible. 
The conclusion is, therefore, that the jet always completely fills 
the nozzle, and that there is no " zonal formation," meaning 
an outer zone moving at a different velocity from the inner one, 

although there is certainly a 
considerable amount of fric- 
tional dragging of the steam 
at the surface. Obviously, of 
course, the statement does not 
hold for absurdly diverging 
forms of nozzles, and in such 
cases the steam leaves the 
walls with apparently much 
loss of velocity as in the ex- 
ample shown by Fig. 25. 

Stodola observed also that 
in any nozzle the pressure 
usually falls in the vicinity 
of the throat to consider- 
ably less than the discharge 
pressure, with a sudden rise 
immediately after the fall. 
This effect is shown by 
pressure curves in Fig. 26, plotted from Stodola' s data taken 
in a divergent nozzle like the one represented at the top of 
this figure. Similar effects, only more pronounced, observed in 
a straight non-expanding nozzle with rounded inlet are shown 
in Fig. 27. Here also a sudden drop below the discharge 
occurs; and, peculiarly, the point of depression progresses along 
the axis of the nozzle as the pressure decreases. Very pro- 
nounced oscillations are set up which extend even into the 
exhaust space for a distance of about one and a half times the 
length of the nozzle. The oscillations are apparently most 
violent for the middle range of pressure, and tend toward a 



a 

b ~ 

_c ^- 



£ 120 



100 



8 60 



12 3 4 5 6 

Inches from the end of the Nozzle 

Fig. 26. Experiments with an Expand- 
ing Nozzle Showing the Effect of Vary- 
ing the Final Pressure. 



NOZZLE DESIGN 



57 



minimum when the lower pressure approaches a perfect 
vacuum. 

In the divergent nozzle, however, there appear to be no internal 
oscillations of pressure after those at the throat have died out. 

The size and most likely also the shape 
of the external space has a considerable 
effect on these oscillations of pressure. 

Jude states in this connection that 
there is a greater loss in velocity, due 
to oscillations or eddies, in a square or 
rectangular nozzle than in a circular 
one. Recent experience with nozzles 
of this type does not bear out this 
statement, except in the case probably 
of square or rectangular nozzles with 
no rounding at the edges. An efficiency 
of 97 per cent, is not unusual for prop- 
erly designed square and rectangular Fig. 27. 
shaped nozzles without any " square 




012 

Inches from Mouth 



Experiments with 
a Non-expanding Nozzle 
Showing the Effect of 

edges ; and circular nozzles have certainly Varying the Final Pressure> 
never given 99 per cent, efficiency. 

Under- and Over-Expansion. The best efficiency of a nozzle is 
obtained when the expansion required is that for which the nozzle 
was designed, or when the expansion ratio for the condition of the 
steam corresponds with the ratio of the areas of the mouth and 
throat of the nozzle. A little under-expansion is far better, how- 
ever, than the same amount of over-expansion, meaning that a 
nozzle that is too small for the required expansion is more efficient 
than one that is correspondingly too large.* Fig. 28 shows a 

* It is a very good method, and one often adopted, to design nozzles so that 
at the rated capacity the nozzles under-expand at least 10 per cent., and maybe 
20 per cent. The loss for these conditions is insignificant, and the nozzles can 
be run for a large overload (with increased pressures) in nearly all types without 
immediately reducing the efficiency very much. This applies especially to tur- 
bines governed by cutting out nozzles in the first stage (see page 277) and with no 
control of the nozzles in the other stages. Under-expansion due to a throttling 
governor is also an important condition affecting the efficiency of nozzles. 



58 



THE STEAM TURBINE 



curve representing average values of nozzle loss used by various 
American and European manufacturers * to determine discharge 
velocities from nozzles under the conditions of under- or over- 
expansion. This curve will be referred to again in connection 
with the design of blades and is very useful to the practical 
designer. 

Non-expanding Nozzles. All the nozzles of Rateau turbines 
and usually also those of the low-pressure stages of Curtis turbines 



10 



<S'§8 

•tf. 



;^2 



25 20 15 10 5 
Percentage Nozzle is too Small 
at .Mouth (Under Expansion) 



+ 



5 10 15 20 25 
Percentage Nozzle is too Large 
at Mouth (Over Expansion) 



Fig. 28. Curve of Nozzle Velocity Loss. 

are made non-expanding; meaning, that they have the same area 
at the throat as at the mouth. For such conditions it has been 
suggested that instead of a series of separate nozzles in a row a 
single long nozzle might be used of which the sides were arcs of 
circles corresponding to the inside and outside pitch diameters 
of the blades. Advantages would be secured both on account of 
cheapness of construction and because a large amount of friction 
against the sides of nozzles would be eliminated by omitting a 
number of nozzle walls. Such a construction has not proved 
desirable, because by this method no well-formed jets are secured 
and the loss from eddies is excessive. The general statement 
may be made that the throat of a well-designed nozzle should have 
a nearly symmetrical shape, as for example a circle, a square, etc., 
rather than such shapes as ellipses and long rectangles. The 

* C. P. Steinmetz, Proc. Am. Soc. Mech. Engineers, May, 1908, page 628. 
A. Jude, The Theory of the Steam Turbine, page 39. 



NOZZLE DESIGN 59 

shape of the mouth is not important. In Curtis turbines an 
approximately rectangular mouth is used because the nozzles are 
placed close together (usually in a nozzle plate like Fig. it 4) in 
order to produce a continuous band of steam; and, of course, 
by using a section that is rectangular rather than circular or 
elliptical, a band of steam of more nearly uniform velocity and 
density is secured. 

Fig. 29 shows a number of designs of non-expanding nozzles 
used by Professor Rateau. The length of such nozzles beyond 
the throat is practically negligible. Curtis non-expanding noz- 




15.19^, r 24 - 20 ". 

Fig. 29. Rateau Non-expanding Nozzles. 

zles are usually made the same length as if expanding and the 
length is determined by the throat area. The Curtis nozzles 
made in Germany are a little shorter than the length calculated 
by formula (9). 

Materials for Nozzles. Nozzles for saturated or slightly super- 
heated steam are usually made of bronze. Gun metal, zinc 
alloys, and delta metal are also frequently used. All these metals 
have unusual resistance for erosion or corrosion from the use of 
wet steam. Because of this property as well as for the reason 
that they are easily worked with hand tools* they are very 
suitable materials for the manufacture of steam turbine nozzles. 
Superheated steam, however, rapidly erodes all these alloys and 
also greatly reduces the tensile strength. For nozzles to be 
used with highly superheated steam, cast iron is generally used, 
and except that it corrodes so readily is a very satisfactory mate- 
rial. Commercial copper (about 98 per cent.) is said to have 
been used with a fair degree of success with high superheats; 
but for such conditions its tensile strength is very low. Steel 
and cupro-nickel (8 Cu + 2 Ni) are also suitable materials, and 
the latter has the advantage of being practically non-corrodible. 

* Nozzles of irregular shapes are usually filed by hand to the exact size- 



CHAPTER IV. 

STEAM TURBINE TYPES AND BLADE DESIGN. 

All the types of both water and steam turbines are commonly 
divided into two general classes, designated by the descriptive 
terms impulse and reaction. Without further explanation, these 
terms, as they are used in turbine practice, would be very mislead- 
ing, because practically all commercial types of steam turbines 
operate by both the impulse and the reaction of steam. Long 
usage, however, has determined the accepted meaning of these 
terms and it is useless now to try to change them. Briefly, the 



Reaction 
Force 




. Impulse 
Force 



Fig. 33. Impulse of a Jet Exerted on a Flat Surface. 



physical phenomena known as impulse and reaction will first be 
described, to be followed by an explanation of the technical sig- 
nificance of these terms as they are used by engineers. 

In all important commercial types of steam turbines the blades 

60 



STEAM TURBINE TYPES AND BLADE DESIGN 



61 



are moved by both the impulse and the reaction of impinging 
steam jets issuing from nozzles (see Fig. 2) or passages essentially 
equivalent to nozzles. According to the older school of scientists, 
who have handed down to us the classification of turbines men- 
tioned above, an impulse is a force acting in a " forward " direc- 
tion, and a reaction is a " backward " force, relative to the 
impulse and equal to it in magnitude. Fig. 33 is a simple con- 
crete illustration of both impulse and reaction. A suspended 
tank filled with water is shown from which a jet issues through a 
nozzle and impinges upon a flat board hung opposite. As the 




Impulse and 
Reaction Forces 



Fig. 34. Impulse of a Jet Exerted on a Curved Surface. 



result of the pressure due to the jet, the board will obviously move 
to the right. As the jet issues from the nozzle it exerts at the 
same time a reaction on the tank causing it to move to the left.* 

* The pressure on the walls of a tank at any point depends on the height of the 
water above that point ("the head") and upon the density of the fluid. When 
a fluid escapes from an opening in the tank there is no resistance at that point to 
pressure, and the unbalanced force exerted on the walls directly opposite will 
tend to move the tank in the direction opposite to that of the escaping jet. The 
greater the "head" and the density the greater will be the velocity of the issuing 
fluid and the reaction on the tank. This explanation of reaction is given here merely 
to show the " nature " of the phenomenon which physicists call " reaction." Many 
engineers think it best to regard as reaction only the force developed as a con- 
sequence of the generation of velocity at the expense of pressure in the moving 
element. 



62 



THE STEAM TURBINE 



Fig* 34 is intended to show the significance of impulse and 
reaction as they are used in regard to turbines. In this case 
water from the tank impinges against the curved surface of a 
wooden block, and before it leaves this surface it is turned back 
upon itself through an angle of 180 degrees. The block is there- 
fore acted on by two forces simultaneously, both tending to move 
it to the right. When the jet first strikes the surface of the block 





Fig. 35. Impulse Wheel with Blades 
of " Single Curvature." 



Fig. 36. Impulse Wheel with 
Blades of "Double Curvature." 



an impulse force tends to move it, and when leaving, there is acting 
in a " backward " direction a reaction equal to the impulse. If the 
jets represented in the two figures have the same velocity and 
density, and frictional losses are neglected, the pressure on the 
block in Fig. 34 will be twice as great as on the board in Fig. 33. 
Fig. 35 shows a nozzle and a blade wheel in which the blades 
have a " single curvature " as compared with the curved surface 
in Fig. 34; that is, the steam in its passage through the blades is 
not " turned back on itself," or in other words, the curvature 
of the blades is less than 90 degrees. If this wheel were held 



STEAM TURBINE TYPES AND BLADE DESIGN 



63 



stationary so that the blades could not move, the steam would 
leave them in a direction nearly parallel to the shaft. The only 
force, therefore, that is effective for moving the blades is the 
impulse. 

Fig- 3°\ on the other hand, shows blades with nearly 180 degrees 
curvature which turn the steam back on itself on leaving. The 
wheel is thus moved first by the impulse force of the steam exerted 
on the blades in the direction of flow, and then by its reaction. A 
blade turning the steam through less than 90 degrees like the one 
in Fig. 35 will exert only about half as much pressure as one turn- 
ing the steam through nearly 
180 degrees like the one in 
Fig. 36. 

A turbine wheel which 
would be called a reaction 
type is shown in Fig. 37. It 
differs from the one in Fig. 36 
chiefly in the blade section B, 
shown at the top of the draw- 
ing. In this type the expan- 
sion of the steam in the noz- 
zle is only partial, and the 
blades are made so that part 
of the expansion occurs in 
them. In the types shown 
in Figs. 35 and 36, on the 
other hand, all the expansion 
is in the nozzles, with no ex- 
pansion at all in the blades.* 

The amount of expansion of the steam in the blades marks, 
therefore, the essential difference between the two important 
types of steam turbines illustrated by Figs. 36 and 37. In im- 
pulse turbines there is no expansion in the blades, while in 

* The turbine wheel illustrated in Fig. 37 is not, however, typical of commercial 
" reaction " types in which there are sometimes as many as 60 to 80 pressure stages. 




Fig. 37. Simple Reaction Wheel. 



64 THE STEAM TURBINE 

reaction turbines " expanding " blades are used, with the result 
that some of the kinetic energy of the steam is changed to 
velocity in flowing through them. 

From the explanation that has preceded it is obvious that both 
of the types represented by the last two figures operate by both 
impulse and reaction. 

Impulse and Reaction of Fluids. The kinetic energy of a 
fluid jet discharging from a nozzle may be regarded as produced 
by a constant impulse force I acting upon a weight W of the 
fluid discharged for one second. During this second the velocity 
has changed from zero to V feet per second and has gone through 
a space of \ V feet. The work done by this force in producing 

the kinetic energy (K foot-pounds per second) is I X -/which 



2 



wv 2 

is equal to K or 

2g 

We have then 



IV WV 2 

2 2g 

i _ wv 

g 



In practice the principal distinguishing feature of reaction turbines is the applica- 
tion of stationary blades for partially expanding the steam. The rest of the ex- 
pansion takes place in the moving blades. 

It is sometimes stated, although inaccurately, that the angles of the moving 
blades may be used as a criterion for distinguishing the two types. According to 
these authorities, the moving blades of impulse turbines are symmetrical like Fig. 36, 
and those of reaction turbines resemble in contour those of Fig. 37. In many 
cases the rule could probably be applied, but there are also many exceptions. There 
are some blades made for Curtis turbines which are not nearly symmetrical, and 
no one would call a Curtis turbine a reaction type. 

The difference between impulse and reaction turbines can be very easily shown 
experimentally by putting a pressure gauge between the nozzle and the wheel. In 
the impulse type, because the expansion is completed in the nozzle, it will be found 
there is no drop in the pressure of the steam in passing through the blades; but in 
the reaction type the gauge will record a higher pressure than that in the casing. 

As these words " impulse " and " reaction " are used at the present time there 
is really little connection between the usual meaning of the words and the ideas 
they are to convey in regard to steam turbines. Actually all commercial steam 
turbines work by impulse and by reaction. A German writer has used instead 
of " impulse " and " reaction " the more accurate words, " gleichdruck " and 
" ungleichdruck," meaning " equal pressure " and " unequ al pressure," which to 
the author seem much more appropriate. 



STEAM TURBINE TYPES AND BLADE DESIGN 65 

In the first principles of physics it was learned that impulse 
and reaction were " equal and opposite," so that if the reaction 

WV 

is represented by R in pounds, then R = I = 

g 

Example. If the vessel shown in Fig. 33 discharges 10 pounds 
of water per second at a velocity of 322 feet per second, what is 
the force I (impulse) pushing the wooden block away from the 
vessel? Ans. 100 pounds.* 

Also what is the force R (reaction) pushing the vessel itself 
toward the left? Ans. 100 pounds. 

Example. If water is discharged against flat blades of a 
water wheel made up of vanes similar to the block shown in 
Fig. 33 at the rate of 32.2 pounds per second at a velocity of 
200 feet per second and is spattered from the wooden blocks 
with a " residual " velocity (leaving the vanes) of 100 feet per 
second, what horsepower is this water wheel capable of devel- 
oping? 

Solution. Calling the " residual " velocity V 2 we have 

W(V 2 ~ V 2 2 ) 32.2(200 2 -IOQ 2 ) .. „ 

K = — - = Q L = 15,000 ft.-lbs. per sec, 

2g 2X32.2 °' ^ 

15,000 , 

or -^ ■ =27.27 horsepower. 

55o 

The maximum theoretical horsepower of the wheel is ■ ? 

2 g x 550 

if the water is discharged at zero velocity. We have (in this 

x ^2.2 X 200 2 20,000 • , . 

case) — ^ = , or 36.36 h.p. 

2 X 32.2X550 550 

The efficiency of the (blades of this) water wheel is therefore 

27.27 

"i—i = -75 or 75 per cent. 

3 6 -3 6 

Example. Steam discharges from a nozzle at the rate of 
3.542 pounds per second with a velocity of 4000 feet per second 
against the blades of a steam turbine and leaves them with a 

* It is assumed that the water leaves the block with practically no velocity, 
that is, all the velocity is absorbed in producing the impulse force. 



66 THE STEAM TURBINE 

velocity of iooo feet per second. Neglecting frictional losses, 
what is the maximum horsepower that this turbine wheel can 
develop? Calculate the efficiency (percentage) of the blades in 
this turbine. Ans. 1500 horsepower; 93.75 per cent. 

Example. The steam discharging from the blades of the 
turbine wheel in the last exercise is finally directed upon the 
blades of a second turbine wheel. Assuming there has been no 
loss of velocity in passing from one turbine wheel to the other 
and that the steam leaves the second one at 100 feet per second, 
calculate the maximum horsepower that could be developed in 
this second turbine wheel and the efficiency of its blades. 

Ans. 99 horsepower; 99 per cent. 

Example. If we consider the two turbine wheels mentioned 
in the two preceding exercises as combined in a single turbine, 
what would be the total horsepower of the turbine and the 
over-all efficiency if frictional and other losses are neglected? 
Ans. 1599 horsepower; 99.94 per cent. 

Suggestion. The same result could have been obtained by 
calculating the total kinetic energy of the combined wheels, 
using V = 4000 feet per second, V2 = 100 feet per second and 
W = 3.542 pounds of steam. 

Example. Remembering that impulse and reaction are equal 
and opposite, what is the force of the reaction against the plate 
supporting the nozzle required to give a velocity of 4000 feet 
per second to a flow of 3.542 pounds of steam per second? 

Ans. 440 pounds. 

Suggestion. Reaction = impulse (/) = 

6 

Example. The area of a nozzle is .322 square inch. How 
many pounds of steam per second having a density of .144 pound 
per cubic foot must be discharged from the nozzle in order to 
exert a pressure of 90 pounds against a plate suitably designed 
to turn away the steam with zero velocity? Ans. .966 pound. 

Suggestion. In this case all the velocity is absorbed in produc- 
ing the pressure (impulse) upon the plate. 



STEAM TURBINE TYPES AND BLADE DESIGN 6j 

Substituting the values given in the example and substituting 
in the equation for impulse, we have 

Jjr . 322 X V X .144 rr 

W = o. ^t — .000322 V, 

144 

r WV .000^22 V 2 T ro 

I = ■ = = .00001 V 2 , 

g 3 2 - 2 

and since the impulse is 90 pounds, we have 

.00001 V 2 = 90 pounds 
V 2 = 9,000,000 
V = 3000 feet per second. 

Substituting this value of V in the equation at the top of the 
page, 

W = .000322 X 3000 = .966 pound per second. 

Example. Steam of the same density as in the preceding 
exercise discharges at the rate of 3478 pounds per hour and pro- 
duces a reaction against the plate into which the nozzle is in- 
serted of 90 pounds. What is the velocity of discharge? 

Ans. 3000 feet per second. 

EXAMPLES OF IMPULSE TURBINES. 

A simple impulse turbine is represented by diagrammatic draw- 
ings in Fig. 38. In the shaded drawings in this figure, " Section 
A " is made by a plane cutting one of the blades and passing 
through the center of the shaft. The other view, " Section B," 
shows a section made by a plane parallel to the shaft and passing 
through the center of one of the nozzles in the turbine. In the 
same figure. Curve I shows the decreasing pressures in the nozzle 
and the constant pressure through the blades. Curve II shows 
similarly the velocity changes. In the nozzle the steam velocity 
increases as the pressure falls, while in the blades the velocity 
of the steam is absorbed in moving the wheel. This simple im- 
pulse turbine represented by these diagrams is typical of the origi- 
nal and simplest De Laval type. (See pages 176 to 183.) These 
turbines have always a single set of nozzles and one row of blades. 




£*<fc 



Velocity Triangles 

Fig. 38. Diagrams of a Single- 
stage Impulse Turbine. 




SECTION A 



*J 



B 



"SECTION B 




(68) 



Velocity Triangles 
Fig. 39. Diagrams of an 
Impulse Turbine with 
Two Velocity Stages. 



STEAM TURBINE TYPES AND BLADE DESIGN 69 

In Figs. 38, 39, 40, and 41 illustrating the important types of 
steam turbines, the direction of the flow of the steam is marked 
by the symbol tn-^ and the motion of the blades by bd— ». The 
moving blades are shown by solid black to distinguish them 
from the stationary blades, which are indicated by cross-hatching. 

A modification of the simple impulse type is shown in Fig. 39. 
The drawings marked "Section A" and "Section B" show a 
turbine with two moving blade wheels and a set of stationary 
" intermediate " blades. The stationary blades are merely guides 
for changing the direction of the steam so that it will enter 
the second set of moving blades at a suitable angle. Two blade 
wheels are used instead o( one in order to make it possible to use 
efficiently a lower peripheral speed for the moving blades. The 
reasons for this statement will be discussed in another part of 
this chapter. The curves at the top of the figure show, graph- 
ically, the relation between pressure and velocity. Curve III 
shows the sudden fall of pressure in the nozzle and the constant 
pressure through the three rows of blades. Curve IV shows first 
the rapid increase in velocity as the pressure falls, and then the 
gradual loss of velocity in the moving blades as it is given up in 
doing work. Velocities represented in Curves II and IV are 
drawn approximately to the same scale. A comparison shows 
that the reduction in velocity of the steam in the first wheel as 
represented in Curve IV is only about half that for the single wheel 
in Curve II. The arrangement of blades represented in Fig. 39 
makes possible comparatively low blade speeds with initially 
high steam velocities. This method of increasing the number of 
rows of blades is often used with three rows of moving blades and 
two "intermediate" (stationary) rows; and even four rows of 
moving blades have been used. Not much advantage, however, 
has been shown from the use of the third and fourth rows of 
moving blades, and this construction has been generally abandoned. 
Turbines of this type are often spoken of as having velocity stages, 
the number of velocity stages being the same as the number of 
rows of moving blades. 

The Curtis turbines, made by the General Electric Company, 



70 THE STEAM TURBINE 

are the best examples of the type illustrated by Fig. 39 with several 
rows of blades following a set of nozzles. In the latest designs 
of the larger sizes of these turbines there are two rows of moving 
blades and one set of " intermediate " blades for each set of 
nozzles, so that the arrangement shown in Fig. 39 is typical of 
these designs.* 

In Fig. 40 another distinct type of steam turbine is illustrated. 
The left-hand half of this figure represents a single impulse wheel 
as in Fig. 38 and the right-hand half is practically a duplicate of 
that on the left. In this construction each of the halves — a 
single nozzle or set of nozzles with the blades following — is called 
a pressure stage, or very commonly it is called simply a stage. 
The difference between the operation of this turbine and the 
single impulse wheel in Fig. 38 is best shown by comparing the 
pressure and the velocity curves at the top of the two figures. In 
Curve I, showing the pressure for the single impulse wheel, the 
steam drops from the boiler pressure to that of the exhaust in a 
single nozzle, that is, in a single stage. In Curve V of Fig. 40 
there is about equal reduction of pressure in each of the two 
nozzles, and the velocity change, as Curve VI shows, is about 
the same for each of the two stages. This figure represents, 
diagrammatically, a number of types that are more complex. 

It should be mentioned here that there are often two or more 
groups of nozzles and blades, each like Fig. 39, in succession (cf. 
Fig. 119). Each of these groups is then called a stage. In other 
words, the first set of nozzles and all the rows of blades up to the 
next nozzle make the first stage, and so on. This last arrangement 
is typical of the Curtis turbines with more than one pressure 
stage and the various Rateau designs. 

* The blades shown in ''Section A" of Fig. 30 have the same height on the 
"entrance" and "exit" sides. It is, however, a very common practice to make the 
"exit" side of the "intermediate" blades of Curtis • turbines a little higher than 
the " entrance " side so as to increase the cross-sectional area and thus allow for 
the lessened velocity, due to friction and eddies, and thereby prevent " choking " 
in the blades. There is therefore a little expansion in these blades. 



STEAM TURBINE TYPES AND BLADE DESIGN 



71 




SECTION A 



if 


*^ 




• 


rs 
^ 


\J 

i 


t 

B 


CO— > 


§ 


^ 

1 



SECTION B 




Velocity Triangles 
Fig. 40. Diagrams of an Impulse Turbine with 
Two Pressure Stages. 



7 2 



THE STEAM TURBINE 




DrumRotor 

SECTION f 



W 4 


s 


*J 




*J 


w 




*s 


^0 




w 




*3# 




*dlP 


*\ 


W., 


^ 


H 

B 


$ 


? 




f|i 



SECTION B 




Velocity Tnaugles 

Fig. 41. Diagrams of a Three-stage Reaction 
Turbine. 



STEAM TURBINE TYPES AND BLADE DESIGN 73 

The Rateau turbine has from 8 to 15 pressure stages, with a 
set of nozzles and a single blade wheel for each. The drop in 
pressure is then, of course, comparatively small in each stage. 

REACTION TURBINES. 

The arrangement of blades in the well-known Parsons turbine 
is illustrated in Fig. 41. This is the typical modern reaction tur- 
bine. There are no nozzles. The steam flows from the boiler 
into the "admission space" of the turbine (see " Section A") 
with practically no velocity. From this space it enters the first set 
of stationary blades, where it expands and attains some velocity 
as the pressure drops. Curves VII and VIII show the change of 
velocity with change of pressure. When the steam leaves the fixed 
blades it enters immediately the first set of moving blades. Here 
it expands again; but at the same time some of the velocity from 
the expansion is taken away, or, in other words, the velocity is 
reduced in moving the blade wheels. The pressure and velocity 
curves show plainly what happens in turbines of this type as the 
steam passes alternately through the fixed and moving blades, 
expanding in every row- till it escapes in the exhaust. There is 
here considerable expansion in the moving blades, and conse- 
quently because the pressure is not the same on both sides of these 
blades it is called a reaction turbine. All the other three types 
(Figs. 38-40) are impulse turbines, because the pressure is practi- 
cally the same on both sides of the moving blades. 

We should observe here that all the possible simple combina- 
tions have been mentioned except the case of expansion only in 
the moving blades and with no expansion in the stationary parts. 
Such an arrangement would be feasible but has probably never 
been used. 

In a reaction turbine any two rows of blades, the first stationary 
and the second moving, make a pressure stage. In a Parsons 
reaction turbine there are sometimes nearly a hundred stages. 

Graphical Diagrams of Steam Velocities. A velocity diagram 
representing graphically the steam velocities in the passages of 
each of four types of turbines shown in Figs. 38-41 is represented 



74 



THE STEAM TURBINE 



at the bottom of each of these figures. These diagrams, in the 
shape of velocity triangles, are represented here with the nozzles 
and blades in their proper order. In practical designing, how- 
ever, this pictorial effect is omitted and only the triangles are 
drawn. The lines of these triangles show by their lengths the 
magnitudes of the blade as well as the steam velocities in the tur- 
bine. As all of these triangles are drawn to the same scale, they 
show how different the velocities are in the four types. In each 
case the blade speed (V&) is taken at about the value that has been 
found by experience to give the best efficiency. Such velocity 
diagrams are used by engineers for determining the best relation 
between the velocity of the blades and the velocity of the steam. 
In order to interpret such diagrams intelligently the significance 
of absolute and relative velocities * of the steam must be clearly 

* This distinction between absolute and relative velocities should probably be 
made plainer for those who are unfamiliar with these terms. A thorough under- 
standing of what is meant by absolute and relative velocities is very necessary to 
work intelligently with the velocity diagrams on which the whole theory of turbine 
practice depends. Suppose a train is just moving out of a station at the rate of 30 feet 
per second, and a man standing in the middle of the track behind the train throws 

a ball with a velocity of 40 feet per 
second through the back door of the 
last car. Then a passenger in the train 
will see the ball moving through the car 
at a velocity of only 10 feet per second. 
In this case the velocity of the ball, or 
40 feet per second, is its absolute velo- 
city with respect to bodies that are not 
moving, and 10 feet per second is the 
relative velocity of the ball in the train. 
In this connection a slightly different 
case should also be considered. Sup- 
pose now the ball is thrown upon a 
boat moving in a stream at a velocity of 
30 feet per second by a man standing 
on the bank at P as represented in Fig. 42. Let us assume the absolute velocity, 
or the velocity with which the ball is thrown, as again 40 feet per second, but that 
now the path of the ball makes an angle of 20 with the direction of the moving boat. 
Then the relative velocity of the ball (V r ) with respect to the direction of the boat is 
shown graphically by a triangle of velocities ABC in the figure, where AC is the 
absolute velocity (VJ of the ball, BC is the velocity of the boat (F&), and AB is 
the relative velocity (V r ) of the ball with respect to that of the boat. 




FiG.42. 



STEAM TURBINE TYPES AND BLADE DESIGN 75 

understood. An absolute velocity of a body is its velocity with 
respect to immovable points on the earth. A relative velocity is 
its velocity with respect to points that are also moving. 

The direction of the line representing the velocity of the 
steam relatively to the blades should be such that the lines of 
flow of the steam enter the blade tangentially to the conven- 
tionally straight portion of the back* of the blade (see Figs. 43, 
49, and 50). If the backs of the blades are made to any other 
angle there will be losses due to impact and eddies. 

EFFICIENCY OF THE BLADES OF IMPULSE TURBINES. 

In the velocity diagram in Fig. 38, the initial velocity of the 
steam entering the nozzle is marked Vi, the velocity in the 
throat is V , and the absolute velocity of the steam as it leaves 
the nozzle and enters the blades is V 2 , making an angle a with the 
direction of motion of the blades. The velocity of the blades 
V 6 , which is the peripheral velocity of the wheel, produces a 
11 relative " velocity of the steam in the blades V r2 . The angle 
£ shows then the theoretical " entrance " angle for the blades 
that the steam may enter without loss of velocity due to shock 
or impact. These angles a and are marked plainly in the draw- 
ing of " Section B." The relative velocity of the steam leaving 
the blades is represented by V r3 . Often the blades for impulse 
turbines are made s)Tnmetrical, so that the angle 7 on the 
" exit " side of the blades is equal to the angle (3 on the " en- 
trance " side. The absolute velocity of the steam leaving the 
blades is found by geometrically subtracting again the blade 
velocity V 6 . The velocity of the blades is always subtracted a 
second time, because the direction of the steam has been reversed 
in passing through them. The steam is discharged with the 
absolute velocity V 3 , which is called commonly the " residual " 
velocity. 

Conditions of Best Efficiency. The condition for the highest 

efficiency of this simple turbine (Fig. 38) will now be discussed. 

The same velocities represented at the bottom of Fig. 38 are 

shown again with the addition of an enlarged section of a blade 

* The " back " of the blade is the side with convex curvature. 



7 6 



THE STEAM TURBINE 



in Fig. 43. The notation is the same as in Figs. 38-41. V 2 and 
V 3 * are the absolute velocities of the steam entering and leaving 
the blade, of which a shaded section is shown. V^ and V r3 are 




2 v 6 
FlG. 43. Velocity Triangles for an Impulse Turbine. 



the corresponding relative velocities of the steam as it passes 
through the blade. Now the energy in the steam is measured, 
of course, in terms of its absolute velocity, and is proportional 
to the square of its velocity.f The energy, then, in a pound of 

steam entering a blade is — and on leaving is — . The energy 
taken away by the blades is, therefore, 



2g 



(V 2 



2£ 

-V 3 2 ). Hereg 



is the acceleration due to gravity (32.2), and for all practical 
purposes is a constant value. Energy converted into work in a 

* Observe that V 2 , V 3 , V4, etc., indicate absolute velocities, and Vr2, V r3 , V r4 , 
etc., are relative velocities. This relation should be of much assistance in reading 
the diagrams. 

The order in the use of subscripts follows the method use for the nozzles in the 
preceding chapters. The subscript 1 is still used to represent the initial condition 
of the steam as it enters the nozzles of an impulse turbine or the first row of sta- 
tionary blades in a reaction turbine, while the subscript o is for the condition at 
the throat of a nozzle. The first " discharge " velocity either from nozzles or 
stationary blades is therefore represented by the subscript 2. 

t See discussion of available (kinetic) energy and velocity, page 24. 



STEAM TURBINE TYPES AND BLADE DESIGN 77 

turbine depends then, theoretically, only on the term (V 2 2 — F 3 2 ). 
This term will have its best value, of course, when F 3 is made 
as small as possible. The best theoretical conditions of blade 
speed and steam velocity are shown in the following discussion: 
In practice it is usual to have given (1) the velocity of the 
steam entering the blades; (2) the " nozzle angle " (the angle at 
which the steam strikes the blades) ; and usually in impulse tur- 
bines still another condition, (3) that the entrance and exit angles 
(0 and 7) are equal. The velocities that must be considered for 
these conditions are shown in Fig. 43. Here V 2 is the absolute 
velocity of the steam entering the blades, the angle a is the " noz- 
zle angle " and shows the inclination of the nozzle to the plane 
of the turbine wheel. V 6 is the peripheral velocity of the blades, 
V r2 and V r3 are the relative velocities of the steam in the 
blades, and V 3 is its absolute velocity leaving the blades. By the 
conditions stated, V 2 and the angle a are known, and we are to 
find the most suitable blade velocity (V 6 ). Also the angle /3 is 
equal to the angle 7, although the value of neither of these angles 
is assumed. The velocities V 2 , V 6 , and V^ will form one triangle 
of velocities, and still another triangle is made with V fe , V r3 , and 
V 3 . The corners of the latter triangle are marked 1, 2, 3, and 
from the geometry of the figure this triangle is obviously equal to 
the triangle 1, 2', 3, marked by cross-hatching. Now, if we as- 
sume there is no loss of velocity due to friction and shock in the 
blades then V r i = V rS , and the triangle i, 2', 3 can then be in- 
verted, and, putting the point 2' at 2, it can be made to join up 
with the triangle o, 2, 3 which shows the initial velocities at the 
upper end of the blade. The base o, 1 of the new triangle o, 1, 3 
is now equal to 2 V b and we can write, by the "Law of Cosines," 
the equation 

V 3 2 = V 2 2 + (2 V 6 ) 2 - 2 V 2 (2 V 6 ) cos a, (1 1) 

or V 2 2 - V z 2 = 4 V 2 V b cos a - 4 V b \ 

V 2 2 - V 3 2 = 4 V 6 (V 2 cos a - V 6 ) . (12) 

In this equation the term (V 2 2 — V z 2 ), which is a measure of 
the energy taken away from the steam, is greatest when 4 V b 



y8 THE STEAM TURBINE 

(V 2 cos a — V b ) has its largest value; * or we get the maximum 
energy taken from the steam when 

V 6 = - V 2 cos a, (13) 

2 

which is the condition when the line 3, 1, or V 3 , is perpendicular to 
V 6 , that is, when the steam leaves the blade perpendicular to the 
plane of the wheel.| 

The condition for which the last set of equations has been 
worked out represents the usual conditions in practice. That is 
the " nozzle angle " is usually assumed (about 20 degrees), and 
the blade angles (3 and 7 are made equal. For this case equation 
(12), above, represents the best blade conditions, with the abso- 
lute velocity of the steam entering the blades (F 2 ) and the 
velocity of the blades (V b ) as the only variables. 

We can express the efficiency of the action of the blades by 
dividing the energy taken away in performing work by the energy 
represented by the velocity of the entering steam; thus, 

Energy taken away for work, or the actual work done = 
V 2 2 -V3 2 - 

2g 

Total energy in the steam, which is a measure of the total work 
possible = — ■• 

2g 

Efficiency actual work done = y^^ = y^ 
total work possible 2 g 2 g V 2 

Now, in equation (12) we have for the best conditions, 
V 2 2 -V 3 2 = 4 V b (V 2 cosa- V b ). 

* If we make the substitution V 2 2 — V3 2 = y, V b = x, K = Vi cos a, then for 
equation (12) we^can write y = 4X (K — x) =4 Kx — 4 x 2 . 
For the maximum value of y, 

g-4C*-..)-a 

x = \ K, or V b = \ Vi cos a. 
t Without the calculus demonstration it is obvious that V2 2 — Vi is largest 
for given values of V2, when Vz is smallest, and this is when the line 3,1 in the 
triangle 0,1,3 is shortest; or, in other words, when the direction of V3 is perpen- 
dicular to the direction of V b . 



STEAM TURBINE TYPES AND BLADE DESIGN 79 
Then substituting this in equation (14), 



Efficiency = 



4 Vb (V2 COS a — V & ) _ 4V6 

v 2 2 = v 2 



(cosa-g). (15) 



If, further, the " nozzle angle " a is 20 degrees, as is so com- 
mon in practice, then 

Efficiency = ^(.940-^). (16) 

Vb 

The only variable left in this equation is the ratio •— , and 

V2 
it follows then that the efficiency of a single row of blades with 
a given nozzle angle and equal entrance and exit angles for the 
blades depends only on the ratio of the velocity of the blades to 
the velocity of the steam discharged from the nozzle. 

Fig. 43 can be used again to determine the best relations be- 
tween blade speed (V&), the absolute velocity of the steam en- 
tering the blades (V 2 ), and the angles /3 and 7. The following 
relations are obvious: 

F 2 2 = VJ +Va 2 -2 V b Vr2 cos (180 - /3) (A) 

Vz 2 = IV + F r 3 2 -2F b F r3 cos7 (B) 

V 2 2 - F3 2 = - 2 V h Vn [cos (180 - 0) - cos 7] (C) 

Equation (C) is obtained by subtracting equation (B) from (A) 
and assuming, as before, that V-^ = F f3 ; that is, neglecting 
blade losses. Putting cos (180 — (3) = — cos 0, we have 

F 2 2 - F 3 2 = 2 V b Vr2 (cos + cos 7). (16a) 

Now the maximum value of V 2 2 — V s 2 , which is the measure of 
the energy taken from the steam, is secured, in terms of /3 and 7, 
when the product VbV^ has its greatest value. It can be 
proved by geometry, or approximately by trial in a triangle 
drawn to scale, that with V 2 and the angle (3 given the product 
of the sides V b and V^ will have a maximum value when V& = V^. 
Now when Vb = V^ the triangle 0,2,3 in Fig. 43 is isosceles and 



80 THE STEAM TURBINE 

each of its acute angles is § /3. The geometry of the figure gives 
then 

Vb cos | jS + F& cos | j8 = F 2 , and 

We have thus obtained a very simple equation for calculating 
the blading of an impulse turbine ; but it must not be overlooked 
that if the entrance and exit angles are not equal, this formula 
must be considerably modified, and the result would not be 
nearly so simple. 

Impulse Force Due to Stream Flow Across Stationary Blades. 
In Fig. 43a a stream of fluid is shown impinging on a blade at 



JEla trance 



Fig. 43a. Stream Lines in Turbine Blade. 

A where the direction of flow is horizontal and parallel to the 
contour of the tip of the blade. At A the stream exerts an im- 
pulse I in the direction of flow, and as it leaves the blade it 
exerts a reaction R, parallel to the direction of flow at the other 
end but opposite to the initial direction of flow. The component 
of R in the direction at which the stream enters the blade (hori- 
zontal) is R cos jS, where /3 is the angle the leaving stream makes 
with its initial direction (horizontal). But since impulse is 
equal to reaction (see page 64), I = R. Consequently the total 
pressure upon the blade due to both impulse and reaction is 

I + Rcos/3 or I(i + cos/3). 



STEAM TURBINE TYPES AND BLADE DESIGN 



8l 



When the stream flow has been turned through 180 degrees 
in its passage over the blade, /3 = o, cos /3 = i, and the total 
pressure is 2 /. It has been shown (page 64) that 

T WV 
g 
and therefore total pressure on the blade is 

2WV 



2I 



g 



Also when j8 = 90 degrees, as is approximately the case in 
Fig- 33 > cos (3 = o, and the total pressure is 

T WV 
X — • 

g 

In Fig. 44 a curve is shown which has been calculated to rep- 
resent equation (16) for varying values of blade speed (F&) and 

















































^ *> 


























































































2 *> 










































N 


s 






































































































































£ 20 




/ 










































1 


/ 












































/ 













































200 400 



600 800 1000 
Blade Speed V^. 



1200 1400 1600 
Ft. per Sec. 



1800 2000 2200 



Fig. 44. 



Curve of Efficiency of an Impulse Turbine with One Row of Blades 
and a Nozzle Angle of 20 Degrees for Varying Blade Speeds. 



with an initial steam velocity ( F 2 ) of 3000 feet per second. The 
increase in efficiency with increased blade velocity should be 
observed, and that the highest efficiency is obtained when the 
blade speed (V&) is about half the velocity of the steam discharged 
from the nozzle (V 2 ). This is a good rough-and-ready rule to 
remember. If, then, the steam velocity is 2500 feet per second, 
the peripheral velocity of the blade wheel, for the highest effi- 
ciency, should be about 1250 feet per second. For mechanical 
reasons it is difficult to construct turbine wheels to run at speeds 



82 THE STEAM TURBINE 

much greater than 500 feet per second, so that many designers 
will generally use low blade speeds to get velocities more suitable 
for commercial application, knowing well that in this respect 
they are sacrificing their highest efficiency. 

In designing blades for turbine wheels the entrance and exit 
angles (J3 and 7) should always be made as nearly as possible of 
the size determined by the velocity diagrams. If the angles are 
made much different, there is a sudden change in the direction of 
the steam instead of a gradual change, with a consequent loss 
due to shock or impact. 

Efficiency of Velocity Stages. An impulse turbine with more 
than one row of moving blades in a single pressure stage (veloc- 
ity stage type) is represented by Fig. 39. The energy taken 
away from the steam for work, as expressed in equation (12), can 
be readily modified to suit this case. We should have observed 
that each time steam passes through a moving blade the blade 
velocity (Vb) is twice taken away (subtracted geometrically) 
in the velocity diagrams. If there are N rows of moving blades, 

V 2 2 - W + 2 = 4 NV 6 (V 2 cos a - NV 6 ).* (12') 



And similarly (compare with equation 15, page 79), 

_] 

V 2 / 



„„ . 4NV6(V 2 cos«-NV b ) 4 NV b / NV 6 \ 

Efficiency = a TTO ' = TT cos a - 

VY V 2 V 



and for a 20-degree nozzle, 

„. 4NV,/ NV & \ , ft . 

Efficiency = ^— L940 - —J. (18) 

Efficiency of a Simple Impulse Turbine for Given Blade Speed. 

In the discussion of the maximum blade efficiency of impulse 
turbines which has preceded, the velocity of the steam entering 

* This can be shown geometrically very easily by the method illustrated at the 
top of Fig. 43 which will be here drawn for three rows of moving blades. As in the 
other figures, V 2 is the velocity of the steam entering the first row of blades and 
Vn = V r z; then in Fig. 45 

F5 2 = V2 2 + (6 Vb) 2 -2F2X6F6 cos a. 

V2 2 - F 5 2 = 12 V b (F 2 cosa - 3 V b ); and F 2 2 - V 2 N +2 = aNVi (V 2 cosa- NV b ), 
if N is the number of rows of moving blades. 



STEAM TURBINE TYPES AND BLADE DESIGN 83 

the blades was assumed to be known and a suitable blade speed 
was determined in terms of the entrance and exit angles, which 
were assumed to be equal. This is the problem which arises 
when a single-stage impulse turbine is to be designed for given 
initial and final pressures. When, however, an impulse turbine 
of more than one stage is to be designed with a fixed blade speed 
(Vb) of say 500 feet per second,* it is desirable to determine the 




Yfc v 6 v 6 v 6 v 6 

Fig. 45 (see footnote). 

pressure drop in the first stage (and probably also in the second 
stage, depending on the action of the valve gear) to obtain the 
highest efficiency in this stage. This is because the best results 
are obtained in most types by getting a larger proportion of 
work from the first stage than from the other stages.f Efficiency, 
therefore, is a more important consideration in this stage than in 
the others. 

We have thus obtained a very simple form for calculating the 
efficiency of an impulse turbine; but it must not be overlooked 

* Many manufacturers have a standard blade speed and all sizes of turbines are 
designed for this standard. The blade speeds of impulse turbines vary from 350 
to 1200 feet per second. The latter figure, it is stated, has been used successfully 
by a European manufacturer. 

f The reason for designing the first stage for the largest amount of work — from 
25 to 50 per cent, more than in any of the other stages — is most apparent in 
turbines operated by " cut-off " governing like the Curtis turbines. This method of 
governing permits a constant standard pressure (presumably that giving the maxi- 
mum efficiency) in the first stage at all loads, while with fluctuating loads the 
pressures will vary considerably in the other stages. But there are also other 
reasons for this method; such as getting down the pressure early so as to reduce 
rotation loss in all stages of the turbine and the injurious effect of high temperatures 
on the blades or buckets. Difficulties as regards packing around the shaft at the 
high-pressure end and " stage leakage " (see pages 101 and 103) are minimized. 



84 THE STEAM TURBINE 

that if the entrance and exit angles are not equal, and in the case 
of velocity stages if the exit angle of the stationary " interme- 
diate " blades is not the same as the angle at which the steam is 
discharged from the preceding blades, these formulas must be 
considerably modified and the result would not be nearly so 
simple. It should be observed also that all losses from friction 
and eddies have been neglected. These more practical con- 
siderations are discussed in connection with the examples of 
actual designs of blades on pages 101 to in. 

EFFICIENCY OF THE BLADES OF REACTION TURBINES. 

As in the case of the impulse turbine, the expressions for energy 
and efficiency will now be derived for the reaction turbine, 
assuming again that there are no losses to be considered. We 
must remember that in the reaction turbine there are no nozzles 
for expanding the steam but that the expansion occurs in both 
the stationary and the moving blades, so that as the steam goes 
through the turbine its velocity is gradually and continually 
changing. 

We shall first consider a reaction turbine (Fig. 46) with only 
two sets of blades. As there are no nozzles, the first set is, of 





v 6 
Fig. 46. Velocity Diagrams for One Stage of a Reaction Turbine. 

course, made stationary. The steam expands in going through 
these stationary blades and attains the velocity V 2 * when it 
reaches the first set of moving blades. The relative velocity with 
which the steam enters the moving blades is Vr2. Now, in these 
blades the steam is again expanded, so that just before it leaves 
the moving blades its relative velocity is V r3 , which is greater thar 
* See note at the bottom of page 76 regarding this notation. 



STEAM TURBINE TYPES AND BLADE DESIGN 85 

Vr2- The absolute velocity at which it is discharged from the 
moving blades is V 3 , and we have the following energy relations: 

the kinetic energy entering the moving blades. 

V2 _ -ir 2 
r3 Vr2 



! = kinetic energy in steam leaving the stationary blades, or 



2g 



= kinetic energy developed in the moving blades. 



— = kinetic energy carried away in the discharged steam. 

2g 

The actual work done on the moving blades is W k = (kinetic 
energy of the steam entering the moving blades) + (kinetic en- 
ergy developed in the moving blades) — (kinetic energy carried 
away), or 

Wl _3e + Sd£3^H. (A) 

2g 2g 2g 

If the steam had left the moving blades with zero velocity, and, 
therefore, no energy had been carried away in the discharged 
steam, the energy available for work would be 

Wam it + J£=Jd tMaA ( B) 

2g 2g 

Effi . = actual work done (A) = V 2 2 +Vr3 2 -V^ 2 -V3 2 * ( , 

° ienCy total work possible (B) V 2 2 +V r3 2 -V r2 2 ' 

In the same way the efficiency can be calculated for any number 
of rows of blades. Equation (19) expresses the efficiency for 
only two rows of blades — one stationary and one moving — or, 
in other words, for one stage. We shall now obtain the efficiency 
for three stages, that is, for six rows of blades. The correspond- 
ing velocity diagram is shown in Fig. 47. 

V 2 

— — = kinetic energy developed in the first stationary blades. 

* Efficiency of a single stage approaches its maximum value as Vz is diminished. 
If Vz could be made zero, the efficiency would be 100 per cent. 



86 



THE STEAM TURBINE 



V rZ 2 - Vr* 2 



2? 



kinetic energy developed in the first moving 



blades. 

JV 

2 g 
blades. 



= kinetic energy in steam leaving the second stationary 



V rb 2 ~ Vr* 2 



2g 



kinetic energy developed in the second moving 



blades. 




Fig. 47. Velocity Diagrams for Three Stages of a Reaction Turbine. 



2 g 

blades. 



1 = kinetic energy in steam leaving the third stationary 



V rl 2 - V r& 2 



2 S 



kinetic energy developed in the third moving 



blades. 



V 9 

— — = kinetic energy carried away in the discharged steam 

2g 

final residual velocity. 



STEAM TURBINE TYPES AND BLADE DESIGN 87 

We observe here that the velocities V a3 and V& 3 are not lost 
but represent velocities that can be effective in the succeeding 
stages. For this reason their energies do not enter the discussion 
of efficiency. The actual work in moving the blades is then 

Wk = ry^ 2 + v* 2 - v^ 2 i + ry^, 2 + ^ - v„n 

L2g 2g J L^g 2g J 

| rVc2 2 | V^-VreH V c3 2 * 
L2g 2g J 2g 

Now, in designing a reaction turbine it is desirable to assume 
that the blade velocities and the corresponding angles of the 
blades are the same and that equal steam velocities are devel- 
oped in each of the three stages, so that 

Va2 = V b2 = V^ V*=Vrt= Vre] VrZ=V r , =V rl\ and VoZ = V c3 . 
l_2g 2g J 2g 

The total energy in the steam available for work in this case is 



L2£ 22 J 



g 

The efficiency is then 

W* = Va2 2 +Vr3 2 -V^-iVa3 2 t , x 

W a V^ + V^ 2 " Vr2 2 K } 

It is clear, then, that in the expression for efficiency the last 
term in the numerator changes its coefficient with the number of 
stages, and we see in what proportion the efficiency is increased 
with the number of stages. 

* In an impulse type it is probable the steam becomes practically " dead " as 
regards velocity before it goes through the next set of nozzles. 

f Observe efficiency approaches maximum value as Vaz (= Vbz — V C 3, residual 
velocity) is reduced, and also as number of stages is increased. With 50 stages 
the coefficient of this term would be -^ instead of |. 



88 THE STEAM TURBINE 

THE MECHANICS AND THERMODYNAMICS OF A STAGE AND A 
GROUP OF STAGES IN REACTION TURBINES. 

The stationary blades of a reaction stage perform the function 
of a ring of converging nozzles, that is, the steam expands 
adiabatically from an initial pressure Pi to a lower pressure at 
exit P 2 and thereby a certain amount of energy E2 is made avail- 
able in the form of the kinetic energy of the jet at exit. The 



p 3 

Fig. 47a. Simple Reaction Blading. 

amount of energy to be thus made available in any row is gen- 
erally settled before the actual designing is begun. 

The relation between the velocity of discharge V 2 and the 
available energy in foot-pounds per pound E2 is as already 
stated, 

2g 

The velocity V& of moving blades is made under normal con- 
ditions slightly less than the velocity of discharge of the steam 
V 2 from the stationary blades. Therefore steam enters the 
moving blades at a comparatively low relative velocity Vr2 and 
at an angle /?, somewhat less than 90 . (Fig. 47b.) 



STEAM TURBINE TYPES AND BLADE DESIGN 



8 9 



A moving row of blades can be conceived as being a ring of 
converging nozzles through which the steam expands adiabati- 
cally from the pressure P 2 mentioned above to a lower discharge 
pressure P 3 * and thereby an additional amount of energy E3 is 





Fig. 47b. Velocity Diagrams for Reaction Blading. 

made available by increasing the relative velocity from W2 at 
entrance to V r3 at the discharge, when assuming no frictional 
losses along the blades, thus 



E3 = 



Vr3 2 - V,2 2 
2g 



Impulse Force due to Stream Flow Across Moving Blades. 

The total "tangential" force F on a moving row of blades 
depends on the amount of steam flowing (W) in pounds per 
second, on the relative entrance and discharge velocities (W2 
and Vrz) and on the entrance and exit angles (/? and 7). The 



impulse of the stream entering the moving blades is 



(Fig. 43), and similarly when leaving is 
„ W 



wv 



wv, 
g 



rZ . 



g 



then 



— \ Vr2 cos |8 + V r3 cos 7 \ . 
g 



cos |3 



(21) 



* The steam pressure P z and density at the exit of the moving row of blades are 
again obtainable directly by means of an entropy-heat chart. See, for instance, 
steam charts by F. O. Ellenwood, published by John Wiley & Sons; and by John 
Morrow, published by Longmans. 



90 THE STEAM TURBINE 

This formula is deduced from first principles of mechanics as 
applied for the case of an impulse blade. The action of the 
steam jet in propelling the moving blades is the same in both 
the impulse and the reaction. It should be noted that the 
formula gives the actual tangential force acting on the blades, 
provided that friction losses along the stationary and moving 
blades have been allowed for in computing the relative velocities 
W2 and V r3 . 

The power transmitted is of course proportional to the product 
of the force F into tangential blade velocity V&, and is equal to 

{ Fr2 COS j(3 + VrZ COS 7 j . 

o 

Efficiency of a pressure stage is therefore written, 

Efficiency . V»(V„ cos .0 + V„ cos T ) ^ ^ 

where E is the total available energy in foot-pounds per pound 
of steam corresponding to the total pressure drop in the stage 
considered. The rate of steam flow W through the stationary 
and moving rows of blades is defined by the following relations 
which form the basis of all calculations for the flow of steam 
through nozzles, thus, 

W = A 2 V 2 d 2 = AaVrfda, 

where A 2 and As are the net minimum passage areas in square 
feet at the discharge from the stationary and movable blades 
respectively, d 2 is the steam density in pounds per cubic 
foot at the discharge from the stationary blades, d 3 is the steam 
density in pounds per cubic foot at the discharge from the 
movable blades, and V 2 and V r 3 are the corresponding steam 
velocities (feet per second) as previously defined. This expres- 
sion gives the effective steam flow; that is, the amount of steam 
flowing between the blades and doing useful work. The leak- 



STEAM TURBINE TYPES AND BLADE DESIGN 91 

age steam at the tip of the blades in the clearance space will be 
considered later. It should be noted that the expression 



(Vr2 COS |8 + V r3 cos 7) V & 



g 



(22a) 



gives the useful output in foot-pounds per pound of steam; and 
that similarly the useful output in B.T.U. per pound of steam is 
given by 

(7,3 cos /3 + F r3 cos 7) Vb 
2X778 



(22b) 



As already stated above, in using these formulas the relative 
velocities V& and V r3 at the entrance and discharge sides of the 
moving row are respectively actual values after proper allowances 




Fig. 47c. Illustration of a Practical Example. 



are made for frictional losses in the stationary and moving rows 
of blades. Proper coefficients to allow for these friction losses 
will be given later. It is sufficient to say here that the kinetic 
energy actually developed in either the stationary or moving 
blades is about .8 to .85 of the available energy, that is 



fE 2 =V 2 y 2g) JEz 



2g 



9 2 ; THE STEAM TURBINE 

where f is a fraction equal to .8 to .85. The relative velocity 
at the entrance to the moving blade V^ is of course computed by 
means of simple trigonometric formula applied to the velocity 
diagram illustrated in Fig. 46. 

Example. Consider a common blade section (Fig. 47b) for 
both stationary and moving blades, having a discharge angle a 
of 20 and an entrance angle (3 of 70 . Referring to Fig. 47c, 
the following relations are obvious, 

7* 7^ V h , 



sin no° sin 20° 


— . • • • • lur stationary uiau 
sin 50 


V rZ V h 


for m ovine* hladp^ 


sin no° sin 50° 




V 2 = 1.23 V„, 


7r2 = 446 Vb, Vr3 = 1-23 Vb 



That is, the ratio of the steam velocity discharging from sta- 
tionary blades to the peripheral blade velocity Vb is 1.23. 

The available energy E 2 in the steam discharged from the 
stationary blades is 

YlA^Ul or E, = hfH, 

2g 2g J2g 

where / is a coefficient allowing for friction losses along the 
blades. The available energy (£2) developed in the moving 
row is 

f P ^ ~ Vr2 2 I,SI 7 6 2 I.3I Vb 2 

J &Z = = or *Lz — — . 

2g 2g J2g 

Total available energy for a stage E = E2 + E s . 

Thus for a peripheral velocity 7& = 350 feet per second and 
/ = .81 (corresponding to a velocity coefficient = .90) the total 
available energy for the stage is 

E = L4lX_35°! = 66oo ft _ lbs lb of steam ( 8 s b.T.U.). 
.81 X 32.16 



STEAM TURBINE TYPES AND BLADE DESIGN 93 
The useful output per moving row of blades is 

F b (T r2 cos/3+ F r3 cosT) _ F b (.i52F & H- 1.158 F») _ 1.31 Vf 
g g g 

= 5000 foot-pounds per pound of steam. 
Efficiency = **- — = 75-6 per cent. 

We shall now consider a number (N) of pressure stages as 
illustrated in Fig. 47, where all the moving blades have the same 
peripheral velocity V and the same blade section (the same 
entrance and exit angles). Also all the stationary blades have a 
common section. This is equivalent to saying that all the 
pressure stages have the same velocity diagram shown in Fig. 46. 
It should be noted that the residual velocity from any moving 
row of blades (except the last) is utilized in the following sta- 
tionary row and all the stages except the first utilize the same 
available energy, E. The pressure and density at the begin- 
ning and end of each row or expansion can again be read directly 
from steam charts assuming adiabatic expansion and following 
a constant entropy line.* This of course presupposes that the 
available energy utilized per stage is known. As already shown 
in the previous example, the available energy per stage for given 
blade angles is in the last analysis proportional to the square of 
the peripheral velocity (V) of the moving blades and is deter- 
mined by it. 

Let Eo be the available energy utilized per pound of steam in 
any of the stationary blades except the first. 

,, V2 2 - Vj 

&2 = ~ > 

*gf 

where / is a coefficient to allow for friction. The available 

* Steam charts like Ellenwood's (John Wiley & Sons) can be used to advantage. 



94 THE STEAM TURBINE 

energy in one pound of steam leaving the first stationary row is 

Ei = — y 

The available energy utilized per pound of steam in any of the 
moving rows is 

£3 = 7 

Available energy in any stage except the first is 

E = Ez + £ 3 = —A V 2 2 - Vi + F r3 2 - V^\. 

2gf 

For any set of blade angles all the steam velocities V 2 , F 3 , etc., 
can be expressed in terms of tangential blade velocity V b of the 
moving rows of blades, by means of simple trigonometric rela- 
tions. The available energy E in terms of blade speed that can 
be utilized per stage is expressed also as follows: 

.. CV b 2 



where C is a constant easily determined for any given set of 
blade angles (see previous example). 

The total available energy for N pressure stages is 

2gf 

As already stated previously the useful output in foot-pounds 
per pound of steam from one row of moving blades is 

( Fr2 COS |8 + VrZ COS 7) V b 

gf 



STEAM TURBINE TYPES AND BLADE DESIGN 95 
Therefore the useful output for a number of N moving blades is 

NVb (Vt* COS ff + VrZ COS 7 ) / n 

~Tf ^ (22C) 

The efficiency of the whole group is the ratio of useful output to 
the total available energy, and (neglecting friction) 

Efficiency = (V„co S/? +V.cos 7 )V> . 

The second term in the denominator shows the influence of the 
residual velocity loss. . 

This formula gives the overall efficiency of the stationary and 
moving blades, taking into account not only the loss due to 
residual velocity from the last row but also the friction loss along 
the moving and stationary blades. 

Rotation between Blade Height and Rotor Diameter. In 
blading of a given mean diameter D (inches) there are four 
factors determining the height of the blades as follows : (i) aver- 
age actual specific volume of the steam v (cubic feet per pound) 
in the row of blades considered, corrected for moisture or super- 
heat as may be necessary; (2) the peripheral speed of the blades 
Vb (feet per second), which is proportional to the revolutions 
per minute; (3) the blade exit angle 7 (degrees); and (4) the 
actual flow of steam through the blades W (pounds per second). 

For a mean diameter D (inches) of the blading the sectional 
area of the annulus which passes steam into the blades is 
3.1416D Xh (square inches), where h is the height of the blades 
(inches) . 

Further, if V a is the velocity of the steam as calculated from 
the available energy, and a constant k is used to represent the 

ratio —f , then the area of the annulus (square inches) can also 
Vb 

Wv X idd. 
be written as — — — r~ ^. Correction of this eauation should 
kVb sin 7 



96 the steam turbine 

also be made for velocity coefficient c from Fig. 51, page 101, and 

... WvX 144 

it becomes -— — r- 32 • 
ckV& sin 7 

The two values for the area of this annulus can then be equated 

thus, 

,m6DXh = ^, but > t , 3--4rtJX N t 

ckv 6 sin 7 60 

and we obtain, 

Wv X 144 X 60 



3.1416D X h 



ck X3.1416D X N Xsiay 

or 

WVX864 ( } 

ckN sin 7 V ; 

when using the approximate value (3.1416) 2 = 10. 

This equation must represent the conditions affecting the 
design throughout the turbine. Data for the determination of 
the right-hand member of the equation are usually available, 
so that it can readily be calculated. Values of blade height h 
and rotor diameter D must then be determined by the " cut 
and try " method of calculation; that is, by choosing one value 
and solving for the other. 

For the greater portion of the reaction blading of a steam 
turbine (usually excluding the last rows of the low-pressure 
section) values of k and 7 are constant and c varies but little, 
so that for practical designing of reaction blading we can write, 

D 2 h = v X a constant. (25) 

This equation is very useful for calculating intermediate 
heights and diameters in a section of reaction blading, after 
having determined either the first row or the last as a basis for 
calculation. 



STEAM TURBINE TYPES AND BLADE DESIGN 



97 



PRACTICAL DESIGNING OF BLADES. 



In designing blades for steam turbines we must determine with 
accuracy, 

(i) The angles for the edges of the blades. 
(2 ) The radial height or length of the blades. 

From the preceding discussion of velocity diagrams and blade 
efficiencies it should be clear how the best angles for the edges are 
obtained. It is first necessary to calculate the velocity resulting 
from adiabatic expansion between the limits of pressures in the 
stage for which the blades are intended. Then velocity diagrams 
must be constructed, varying the blade angles if the blade speed is 
assumed till the best efficiency is obtained. This will be when 
the steam leaves the last blades nearly at right angles to the 
plane of the wheel; that is, when the absolute velocity of the steam 
leaving the blades is, in the diagram, nearly perpendicular to the 
line showing the blade velocity (see pages 76 and 78). 

Design of Blades for Impulse Turbines. We shall continue with 
the discussion of the design of blades for an impulse turbine with 
nozzles and with a single row 
of blades, assuming now that 
the entrance and exit angles 
(/? and y) have been deter- 
mined. We shall assume also 
that the total area of the noz- 
zles at their largest section has 
been calculated as it has been 
explained on pages 44 to 49. 

To avoid losses by eddies, 
nozzles are often arranged in 
groups placed symmetrically 
with respect to the periphery fig. 48. 
of the blade wheel. Usually 
the nozzles would be arranged 

in two groups diametrically opposite in a circular plate, called a 
diaphragm, as in Fig. 48. We shall assume that each nozzle 




Diagram Showing Location of 
Nozzles in a Diaphragm. 



98 THE STEAM TURBINE 

group covers one-fourth of the circumference of the blade wheel. 
Then if the blades in the wheel were removed so that they 
could not obstruct the flow of steam, the area through which 
the steam can pass is approximately J 7rDh for each nozzle 
group, where D is the mean diameter of the blade wheel and h is 
the height of the opening from which the blades have been re- 
moved. When, however, there are blades on the wheel the 
height h must be increased, because the effective area for the 
passage of steam is reduced. 

Fig. 49 shows two views of a small segment of a blade wheel. 
The pitch of the blades is marked p and the blade angle is 0. If 
there are no blades, the area for the passage of steam in a length 
p is approximately p X h. With the blades in the wheel the 
area is only h X p sin 0.* It follows then, when we have blades 
under the nozzle groups, that the effective area under each group 
is h X i 7rD sin /?. If we call A the total area of the nozzles at 
the largest cross-section (mouth) we can write 

A = \ TrDh sin/3 + \ TrDh sin /?. 
A = I TrDh sin 0. 

h = J* A a (26) 

ttD sin 

For steam at very high velocity the height of the blades as 
calculated will be too small for practical working conditions; so 
that blades less than .25 inch high are not often made. This 
minimum height is determined most on account of mechanical 
difficulties; but steam leakage through the clearance outside the 
blades also becomes excessive when very small blades are used. 

In impulse turbines with only a few stages no effort is made to 
make use of the velocity, as such, of the steam leaving the last 
blades of a stage. This means some loss; and more experimental 
work might well be done with the object of showing how this loss 
can be turned to better account. 

Fig. 50 shows how impulse turbine blades are laid out. The 

* It is assumed in such calculations that the thickness of the edges of the blades 
is practically negligible. 



STEAM TURBINE TYPES AND BLADE DESIGN 



99 






r 



SECTION AT CENTER LINE 




designer must first decide how wide his blades shall be. For 
turbines of less than ioo horsepower the width of the blades is 
often made about i inch, increas- 
ing this dimension to about 
1.5 inches in turbines of 1,000 
horsepower. The pitch or cir- 
cumferential distance between 
consecutive blades is made about 
.5 to .6 of the axial width.* In 
Fig. 50 the pitch is shown by 
the distance between the points 1 
and 2. Hence, when a drawing FlG 49 Diagram Illustrating the De _ 

of blades is Started these points sign of Blades for Impulse Turbines. 

should first be located. At any 

point between 1 and 2 mark another point 3 and through it draw 
a line $3', making an angle with the horizontal equal to the blade 

angle 7 on that side. Draw 
through 2 a line perpendic- 
ular to the last line and 
intersecting the center line 
of the blades. Mark this 
point on the center line 5. 
Draw through 1 a line par- 
allel to 3 3' to intersect 2 5, 
at 4. With 5 as a center 
draw an arc tangent to 
1 4, which completes the 
lower half of the convex 
side of the blade. With 
the same center the concave side of the next blade is drawn with 
the arc passing through 2. The arrows in the figure show 

* The most efficient blade pitch appears to be between the limits of J inch and 
1 inch. Between these two values the efficiency of blades made according to 
conventional designs is practically constant. The usual blade pitches are f, f, 
and I inch. Even for very small turbines not much less than £-inch pitch should 
be used. Designers usually make the pitch and axial width increase a little with 
the height of the blades. 




Fig. 50. Diagram Illustrating the Method 
for Laving Out the Blades of an Impulse 
Turbine. 



IOO THE STEAM TURBINE 

plainly the center for these arcs. This construction makes the 
" perpendicular " width of the steam passage nearly constant. 

Blade Velocity Losses. Various attempts have been made by 
experimenters to determine the velocity losses in blades with 
stationary apparatus, usually by some method of measuring the 
reaction somewhat in the same way as the coefficients given in 
Fig. 28 were obtained for nozzles. Such results, however, are 
not satisfactory for application to designing. Frictional, eddy, 
and impact losses in moving blades are certainly very different 
from what they are in stationary blades. Apparently there are 
only two ways to get good data regarding these losses. Either 
the velocity must be measured between the blades of an operat- 
ing turbine with a Pitot tube or similar device, or they must 
be determined by the " cut and try " method that has been 
generally followed by designers. The latter method is certainly 
expensive and a slow one for obtaining results. It seems, there- 
fore, that more work should be done along the line of the former 
-method by the application of the Pitot tube. In the latest designs 
of steam turbines there is no difficulty about getting sufficient 
space for a pressure tube between the blades, as the axial clear- 
ance in large turbines is often as much as ^ inch. 

Fig. 51 shows values of the velocity coefficients to be applied 
in designing steam turbine blades. Curve A is for blades re- 
ceiving steam from nozzles. Curve B is for stationary blades, 
or for moving blades receiving steam from stationary blades. 
Values of both curves vary with the relative velocity of the steam 
in the blades. The true velocity in the blades is found by mul- 
tiplying the theoretical relative velocity by the coefficient from 
the curves.* The values given by these curves may be taken 
as fairly representative for all the well-known commercial types 

* Values given by these curves agree well with the determinations made by 
Rateau, Stevens, and Hobart from the analysis of the losses in operating turbines. 
Hobart calculated that the blade frictional and eddy losses in a 275-horsepower 
De Laval turbine are 17 per cent, of the steam velocity which in this case is about 
4000 feet per second. He states also that generally in turbines of this type this 
loss is about 15 per cent, of the relative velocity in the blades. It is stated that 
designers of Rateau turbines assume a blade velocity efficiency of 96 per cent, at 
relative velocities of about 600 feet per second. Obviously near zero velocity the 



STEAM TURBINE TYPES AND BLADE DESIGN 



IOI 



























Ss 












































































































S A 














































N B 























































































500 1000 1500 2000 2500 

Relative Velocity of Steam 

in Blades Ft. per Sec. 



in which the blades have smooth surfaces and the entrance 
edges are made comparatively sharp and at a true angle. These 
curves are intended to be read 
for only two significant figures. 

The initial steam velocities 
in turbines of the Parsons type 
vary from 300 to 700 feet per 
second, in Rateau turbines from 
500 to 1500 feet per second, in 
Curtis turbines from 1500 to 
3000 feet per second, and in some 
single stage types (De Laval) 
from 2 500 to 4500 feet per second. 

The efficiency of energy con- Fig ^ BIade Vdodty Coeffidents . 

Version in the blades of Steam Curve A for moving blades following 
turbines varies from 60 tO 70 per nozzles. Curve B for stationary blades 

cent.* for sizes from 300 to 3000 or for moy '^ blades followin S sta " 

, ill tionary blades. 

kilowatts,! and is roughly about 

50 per cent, for smaller sizes of impulse turbines down to about 
10 kilowatts. Still smaller sizes may have efficiencies less than 
50 per cent., depending largely on the type of construction. For 
any size of turbine, then, the difference between 100 per cent, 
and this efficiency of energy conversion is the loss due to disk and 
blade rotation, leakage, residual velocity, and radiation. 

In a well-designed turbine of say 300 to 500 kilowatts' capacity, 

loss is practically zero. Designers of Parsons and Curtis turbines must use some- 
what larger coefficients (cf. Curve B) for combinations of stationary and moving 
blades, because stationary blades are not as efficient as nozzles. The data for 
these curves were obtained by measuring with modified Pitot tube apparatus the 
velocity of steam discharged from stationary blades of various designs. The 
steam was directed upon the blades from calibrated nozzles. 

* In stating this efficiency it is assumed that adequate provision is made in 
these turbines to prevent leakage: in impulse turbines, between the diaphragms 
and the shaft " stage leakage "; and, in reaction turbines, over the ends of the blades 
through the radial clearance. This leakage loss is as much as 10 to 15 per cent, 
in some good commercial turbines. It should be reduced, however, to not more 
than 5 per cent. 

t A well-known engineer states that the energy efficiency of the 9000 to 12,000 
kilowatt turbines installed in New York and Chicago is as high as 80 per cent. On 
a conservative basis, however, about 70 per cent, can be assumed for 5000-kilowatt 
sizes and 75 per cent, for 10,000-kilowatt sizes. 



102 THE STEAM TURBINE 

the sum of the losses due to friction, disk and blade rotation or 
" windage," leakage, residual velocity, and radiation losses is, 
therefore, about 40 per cent. For turbines of from 2000 to 5000 
kilowatts these losses amount to only about 30 per cent. But 
these are not all actual losses. The energy equivalent of each of 
these losses, except that due to radiation, which is very small, is 
immediately converted into heat, and is partially regained in re- 
heating the steam. The sum of these losses actually reheating 
the steam, expressed as a percentage of the total available energy, 
is called the reheating factor. 

DESIGN OF BLADES FOR AN IMPULSE TURBINE. 

Blades are to be designed for a 300-kilowatt turbine to operate 
with steam at 50 F. superheat, at an initial pressure of 165 
pounds per square inch absolute, and exhausting at 1 pound per 
square inch absolute (approximately 28 inches vacuum). Blade 
speed Vb is 500 feet per second at the rated speed 3600 r.p.m. 
It is assumed that the nozzle will be correctly designed for the 
pressure, so that the nozzle velocity loss is 3 per cent. Governing 
is to be accomplished by the method of " cutting out nozzles " 
in the first stage (see page 277). By this method a practically 
constant steam pressure is maintained in the nozzles of the first 
stage from fight load to overload, and the velocities in this stage 
are at all loads approximately those giving the best blade effi- 
ciencies. In the other stages, however, where the number of 
nozzles open is not controlled by the governor, the velocities are 
variable. For this reason a large pressure drop is to be used 
for this stage,* and to utilize the resulting high velocity efficiently 
there are to be two velocity stages in this pressure stage. A 
reasonable value for the first stage pressure is about 35 pounds 
per square inch absolute. The other stages are to be designed 
for highest efficiency with a single blade wheel in each pressure 
stage. Such a design will be a compound type — the first stage 
resembling the Curtis, and the other stages the Rateau turbines. 

The energy available from adiabatic expansion in the first stage 

* See footnote on page 83. 



STEAM TURBINE TYPES AND BLADE DESIGN 103 

nozzles (as read from the entropy-heat chart) from 165 pounds 
per square inch absolute and 50 F. superheat to 35 pounds per 
square inch absolute is 122 B.T.U. Disk and blade rotation 
losses, leakage between the stages through the clearance between 
the shaft and the diaphragm, and residual velocity of the steam 
leaving the blades amount to 40 per cent. ; and it is assumed that 
all this energy appears again as heat produced by disk and blade 
friction, steam impact, eddies, and throttling. There is then 40 
per cent, of 122 B.T.U. , or nearly 49 B.T.U. , going to reheating 
the steam. This reheating occurs, of course, at the pressure in 
the first stage (35 pounds). As the result of reheating, the 
quality of the steam in the first stage is changed from .932 to 
.985, and the total heat of the steam going to the nozzles of the 
next stage is increased from 1103 to 1152 B.T.U. Fig. 52 shows 




1225 1152 113S. 1103 B.T.U.gcaie 

Fig. 52. Entropy-Heat Diagram for the Design of an Impulse'Turbine. 

graphically this reheating effect. It serves also to show the 
complete energy distribution as required for this design. In 
each stage, as in the first, the reheating is assumed to be 40 
per cent. 

Since all the stages after the first are to be of the single wheel 
impulse type, it is obvious that a large number of stages will be 
needed in order to absorb the velocity of the steam efficiently. 
In a stage of the single wheel type the velocity of the steam should 



104 THE STEAM TURBINE 

not be greater than twice the blade speed. Equation (17) 
shows the relation between the steam and blade velocities for 
the highest efficiency, and this equation can be used for determin- 
ing quite accurately the best energy distribution. Because a 
designing coefficient (C) must be inserted to correct for the velocity 
loss in the blades, this equation will now be written 

v> = -^r-o- ("TO 

2 COS J p 

Now the velocity coefficient for single blade wheels is about 
.95.* The angle ft is usually about 40 degrees. Blade speed, 
Vb, is 500 feet per second. Then 

T7 2 X 500 X cos 20 , , 

V 2 = - = 080 feet per second. 

•95 
But from equation (2) we have the relation that the available 
energy (E a ) in terms of velocity is 

V223.7/ 

&~(^Y-i 9 .6B.T.U. 

V223.7/ 

It is shown then that the required energy per stage is between 
19.5 and 20 B.T.U. per stage. The energy distribution with 
reheating (40 per cent.) was calculated with the help of the chart 
for 19.5 and for 19.8 B.T.U. per stage, and it was found possible 
to get almost exactly equal energy distribution with 12 stages 
each of 19.8 B.T.U. between 35 pounds pressure (quality .985) 
and the exhaust pressure 1.0 pound. This distribution is shown 
in Fig. 52. The quality of the steam in each stage is recorded, 
so that the disk and blade friction can be calculated later from 
the formulas in Chapter V. 

* See Fig. 51. To determine an approximate value for this coefficient the 
probable relative velocity must be estimated. If a very large error were made 
in assuming this coefficient it would be discovered as soon as the velocity diagrams 
are made, as the relative velocity and the coefficients can then be accurately deter- 
mined. As a first approximation in this calculation the general principle in impulse 
blading may be assumed as stated on page 81, that for best efficiency in a single 
row of blades, the absolute velocity of the steam entering the blades is twice 
the blade speed F&. 



STEAM TURBINE TYPES AND BLADE DESIGN 



I05 



Velocity of the steam discharged from the first stage nozzles 



is 



V 2 = .97 X 223.7 v/ i22 = 2398 feet per second, 

and that from each of the other stages is 

y% = -97 X 22 3-7 V19.8 = 965 feet per second. 

The velocity coefficients given in Fig. 51 have been used to 
lay out the triangles in Figs. 53 and 54. The application can 

be best illustrated by the 
triangles in Fig. 53, show- 
ing the velocities of the 
first stage. 

For constructing the 
triangles in Fig. 53, V 2 is 
laid off inclined 20 de- 
grees (the nozzle angle) 
to the horizontal and 
made to scale 2398 feet. 



^\>^ 






**\>\ 




B = .25K/ ^^^ 


«^A=20° 


C=-25M° A ^*^S 


v 6 






^s^ y^\p= 


35° 
20"" 




F = 357 N V 




G 5" 35V/ 


v b 
v 5 






Fig. 53. Velocity Triangles for Two Velocity 
Stages in One Pressure Stage. 



Fig. 54. Velocity Triangles 
for a Simple Impulse Wheel. 



To the same scale the blade speed (F&) is laid off for 500 
feet, making the relative velocity (VV 2 ) in the first row of 
blades 1938 feet per second, and the entrance angle (B) of 
these blades is found to be 25^- degrees. The entrance and 
discharge angles will be made equal, so that the angle C is also 
2 5i degrees, determining the slope of the relative velocity (F r3 ). 
The velocity coefficient taken from curve A in Fig. 51 for a 
relative velocity of 1938 feet is .88, so that V r3 = 1938 X .88 or 
1705 feet. Vb is again laid off in a horizontal direction, and the 
absolute velocity of the steam discharged from the first row of 



106 THE STEAM TURBINE 

blades (F 3 ) as read by the scale is 1270 feet, and the true dis- 
charge angle (D) is 35 degrees. In order that the steam may- 
enter the stationary intermediate blades without shock, the 
entrance angle of these blades must be also 35 degrees, and the 
discharge angle (E) will be made 20 degrees, the same as the nozzle 
angle. The velocity coefficient is now read from curve B in 
Fig. 51 for 1270 feet,* which is .87, and V 4 is laid off for 
1270 X .87 = 1 105 feet. Completing the triangles, V n is 662, 
and the entrance and discharge angles F and G are each 35 
degrees. The velocity coefficient (read from curve B) is .93, so 
that Vr 5 is 615 feet and the final discharge velocity (F 5 ) is 355 
feet. 

Velocities and blade angles are determined in the same way 
(by applying a velocity coefficient) for the 12 single wheel stages 
as shown in Fig. 54. 

Data and results of these velocity triangles are tabulated 
below for convenient reference: 

Blade Angles and Velocities of First Stage. 

First row (moving) : entrance and discharge angles 25 \ degrees. 

Intermediate (stationary): entrance angle 35 degrees; discharge 
angle 20 degrees. 

Second row (moving): entrance and discharge angles 35 
degrees. 

Vb = 500 V r3 = 1938 X .88 = I705 V r4 = 662 

V 2 = 2398 V 3 = 1270 V rh = 662 X .93 = 6l5 

V T2 = 1938 V 4 = I27O X .87 = IIO5 V, = 355. 

Blade Angles and Velocities of Second to Thirteenth Stages. 

Single row (moving) : entrance and discharge angles 39 J degrees. 

V h = 500 VrZ = 525 X .96 = 504. 

V 2 = 9 6 5 ^3 = 339- 

Vr2 = 525 

A slightly higher efficiency could have been obtained if the first 
stage pressure had not been assumed but had been determined 

* In stationary blades the absolute and relative velocities are equal. 



STEAM TURBINE TYPES AND BLADE DESIGN 107 

by a "cut and try" method to get the highest efficiency. If the 
energy for this stage had been a little less, the efficiency would 
have been increased — although an insignificant amount. It is 
a good rule to remember that with a given blade speed, whenever 
the line representing the residual velocity slopes toward either 
side of the vertical, the minimum residual velocity has not been 
found. A higher efficiency could have been obtained also by 
reducing the discharge angle of the intermediate blades. This 
angle is usually made about the same as the nozzle angle (about 
20 degrees in most types). If it is made less than 20 degrees, 
although the apparent efficiency will be increased, there will be 
probably a greater loss than gain on account of the steam spilling 
over the blades. 

Stage Efficiencies. Nozzle efficiency is assumed to be 97 per 
cent., on the basis of the velocity developed. Efficiency of the 
energy conversion in the blades can be calculated from the 
results given by the velocity triangles in Figs. 53 and 54. 

In the first stage the velocity absorbed in moving the turbine 
is the initial velocity (V 2 ) less the residual velocity, (V 5 ), and 
the velocity losses in the blades are (V n — V r3 ) and (V r4 — F r5 ). 
Then the energy absorbed in the first stage,* or 

Work Done = [F 2 2 - (F^ 2 - F r3 2 ) - (F 3 2 - F 4 2 ) 
~(V ri 2 - V r5 2 )-V?? + 2g]. 

-m j x-rc • Work Done 

Blade Efficiency = —r 

Work Possible 

V 2 2 ~ Vr 2 2 + Vr? ~ V 3 2 + VI - Vr? + F r5 2 - F, 2 

V 2 2 
Blade Efficiency (first stage) = 

(2, S q8) 2 -(1Q38) 2 + (170O 2 - (1270) 2 + (nop 2 - (66 2 ) 2 -4- (61. Q 2 - fo.Q* 

( 239 8) 2 

Blade Efficiency = 75.3 per cent. 

Nozzle and blade efficiency of the first stage is therefore 
75-3 x V97 * = 74.2 per cent. 

* When writing efficiency equations, it must be remembered that efficiencies are 
proportional to the available energies and to the square of the velocities. 



108 THE STEAM TURBINE 

Similarly for the second stage (also third to thirteenth stages') 
we have, 

Blade Efficiency = V * ~ V * ~^ n% ~ Vr * ] 

Blade Efficiency = <**>' ~ ^~^ & + <•**>' 

Blade Efficiency = 85.3 per cent. 

Nozzle and blade efficiency of the last twelve stages is therefore 
85.3 X V.97 = 84.0 per cent. 

The combined or " total " nozzle and blade efficiency of the 
turbine, prorated according to stage energy, is, then, 

74.2 X 122 +84.0 X 19 8 X 12 = go 8 ^ * 
122 + (19.8 X 12) 

Besides the nozzle and blade losses, there are bearing losses, 
including the friction of the gland or stuffing-box on the shaft 
and the power for the governor and oil pumps, amounting to 
about 2 per cent in a turbine of this size, f The radiation loss is 
about 1 per cent, the loss due to leakage of steam along the shaft 
between the stages should not be more than 7 per cent., and the 

* Although velocity stages do not give as high net blade efficiency, the adoption 
of this type for the first stage makes it possible, because of the large available energy 
required for this stage by this method, to make the turbine very economical at light 
loads. By providing a suitable valve gear the number of nozzles open in the first 
stage can be controlled by the governor. (See pages 277-290.) 

f Bearing loss in turbines is usually very small. According to Lasche of the 
Allgemeine Electricitat Gesellschaft, Berlin, the friction coefficient (J) is 

/= *+ (txp), 

where / is the temperature of the bearing in degrees C. and p is the pressure in 
kilograms per square centimeter. The rotor of a 1000-kilowatt Parsons turbine 
weighs about 3000 pounds, and the disks and shaft of an impulse turbine would 
probably weigh less. 

Langen in the Zeitsch. fiir das Gesamte Turbinenwesen (Oct. 19, 1907) states 
that the bearing (journal) friction of a well-designed Parsons turbine is about 
.2 per cent., and that the total friction loss including governor and oil pump rarely 
exceeds 1 per cent. 

Stodola's tests of a Zoelly turbine, with, of course, a much shorter casing than 
that of a Parsons type, show the radiation loss from the casing to be about .7 per 
cent. 



STEAM TURBINE TYPES AND BLADE DESIGN 109 

actual net loss of heat (" available ") due to rotation of disks and 
blades will be about 10 per cent.* The sum of the bearing, 
radiation, leakage, and rotation losses is then about 20 per cent., 
and the efficiency of the turbine as measured by work done at 
the shaft is about 81 per cent, (from page 108) less 20 per cent., 
or about 61 per cent. 

The theoretical steam consumption (water rate) of a perfect 
engine operating with steam at the same initial pressure, super- 
heat, and exhaust pressure is 10.24 pounds per kilowatt-hour. f 
Since the shaft efficiency is 61 per cent., the equivalent steam 
consumption per shaft kilowatt-hour developed in the blades is 
10.24 -7- .61, or 16.80 pounds. Generator efficiency might be 
assumed to be about 92 per cent, for a good design suitable for 
this high speed relatively to the size, and the steam consumption 
per kilowatt-hour "at the switchboard" would be about 16.80 -*-. 
.92 = 18.26 pounds.J 

The energy efficiency, neglecting losses, of each stage with a 
single row of blades can be expressed approximately by equation 
(16), thus, 

Efficiency = - — 7^ — (.940 — ^— -J =87 (nearly). 

The nozzles for this turbine must be designed to discharge 
at full load 300 X 18.26 pounds or 5478 pounds per hour at 
50 F. superheat. Total " throat area " of the nozzles (A ) can 
be calculated by equation (5') for superheated steam, where 

* The actual rotation loss for this design can be calculated by the formulas 
given in the following chapter. But a large part of the "total" loss as calculated 
becomes again available as the result of reheating. The mean pitch diameter of 
the blades is 

500 X 60 

or 2.65 feet. 

3.1416 X 3600 

t A kilowatt-hour is equivalent to 2,654,400 foot-pounds or 3412 B.T.U. per 
hour (44,240 foot-pounds per minute). In this case the total available energy- 
taken as one expansion is (1225 — 892) 333 B.T.U. per pound of steam, and the 
theoretical steam consumption is 3412 -=- 333, or 10.24 pounds. 

t Guaranteed steam consumption would be about 10 per cent, more than the 
estimated water rate. It is the usual practice of manufacturers of steam turbines 
and engines to add a percentage of about this value to allow for possible defective 
workmanship in construction. 



no 



THE STEAM TURBINE 



D = 50 degrees, F = 5478 ■*■ 3600, or 1.52 pounds per second, 
and Pi = 165 pounds. Then 

A _ 60F (1 + .00065D) _ 60X1.52 X 1.0325 _ ~ c ^ n • 

The valve gear will be designed to open 8 nozzles for the first 
stage at full load, with provision for opening 4 more at overload, 
so that 50 per cent, overload can be carried efficiently by the 
turbine. These nozzles will all be of the same size. Each first 
stage nozzle will have a " throat area" of .665 4- 8, or .083 square 
inch. It will be assumed that the section of the nozzle at the 
throat is approximately square (with rounded corners) and that 
its width (in the radial direction with respect to the blade disk) is 
constant from throat to mouth, or is V.083, or .288 inch. 

A calculation should now be made to determine the height 
of the blades to give sufficient area for the passage of the 
steam. For this purpose the length of the nozzles at their 
mouths must also be calculated. It is obvious that a nozzle 

cannot be de- 
signed to be cut 
off at the end 
of the expanding 
portion at right 
angles to its ax- 
is; but an exten- 
sion or " tail " is 
necessary to di- 




FiG. 55. Details of the Nozzle Mouth. 



rect the steam upon the blades. To avoid spreading the jet 
and making the expansion ratio uncertain, this " tail " is often 
made non-expanding, so that its wall is parallel to the axis. 
The varying dimensions of the nozzles for this design can be 
determined then from the expansion ratio, which, according to 
the curve in Fig. 21,* is approximately 1.52 for the expansion 
in the first stage nozzles. Area at the jnouth is .083 X 1.52 = 
.1261 square inch, but as one dimension is constant, the longer 

* Although the " expansion line " in this figure was calculated for dry saturated 
and wet steam it can be used with sufficient accuracy for cases in which the initial 
superheat is less than ioo° F. 



STEAM TURBINE TYPES AND BLADE DESIGN in 

side of the rectangular mouth is .1261 -=- .288, or .438 inch 
(marked y in Fig. 55) . By the geometry of the figure it is obvious 
that the length z = y -£■ sin 20 degrees, if the nozzle angle is 
20 degrees as it is generally made. Then the length of the 
nozzle mouth opposite the blades is z = .438 ■*- .342, or 1.28 
inches. 

Sufficient area must be provided in the blades to pass the 
steam from the nozzles. The pressure of the steam in the 
blades is 35 pounds per square inch absolute, of which the spe- 
cific volume (dry) is 11.89 cubic feet. The weight of steam 
flowing per second is 1.52 pounds and the volume (x = .985) pass- 
ing through the blades per second is approximately 11.89 X .985 
X 1.52 = 17.8 cubic feet. This volume and the velocity of the 
steam determine the necessary height (h') of the blades. Net 
area in square feet between the blades for passing the steam 
from eight nozzles may be written as (8 X 1.28 X h f X sin 25J ) 
-7- 144 (see page 98) . This area multiplied by the relative veloc- 
ity of the steam in the blades is another expression for the vol- 
ume. Now the velocity of the steam in the blades, on account of 
the frictional losses, is variable. Obviously the blade area must 
be made large enough to pass the steam at its lowest relative 
velocity; that is, when it is discharging from the blades. The 
height of the first row of blades in the turbine (F r3 = 1705) is 

7 , 17.8 X 144 ( 1 \ 

h = — : 777 ■ = .xa (nearly). 

8 X 1.28 Xsin2 5 i°X 1705 

After the angles and height of the blades * have been deter- 
mined they can be laid out according to the diagram in Fig. 50. 

The effective blade area should be calculated in the same way 
for every row of blades in each of the stages, as it is very impor- 
tant that the blades are provided with sufficient area. 

DESIGN OF BLADES FOR A REACTION TURBINE. 

The rotor of a " simple " Parsons turbine, except for marine serv- 
ices, is made commonly in three sections of different diameters. 

* The other two rows of blades in this same pressure stage are calculated from 
equation above using respectively F 4 = 1105 (stationary blades) and V r b = 615. 



112 THE STEAM TURBINE 

At the high-pressure end of a turbine of this standard type a 
section of small diameter is used, and the intermediate and low- 
pressure sections are made relatively larger to allow for the in- 
creased volume of the steam as it expands in the blades. In 
an impulse turbine, on the other hand, the blade wheels in the 
several stages are usually of the same diameter. This radical 
difference in the type of construction between impulse and 
reaction turbines becomes necessary because of the difference 
in the manner of admitting steam, which results from admitting 
in the reaction turbine high-pressure steam around the whole 
periphery of the rotating part, while in the impulse turbine 
the admission steam is discharged through nozzles, which in 
the high-pressure stages occupy usually only a small part of the 
periphery of the blade disks. 

The passages along the blades in the Parsons reaction type 
are so proportioned that the pressure drop along the stationary 
blades produces a steam velocity which is enough greater than 
the blade speed to overtake the following moving blades. Hence 
a further pressure drop and steam expansion beyond that occur- 
ring in the stationary blades is necessary in the moving blades 
in order to produce the needed impulse to perform work. The 
steam passages along the moving blades are shaped, therefore, 
so as to permit this expansion. 

Because of the expansion of the steam in the moving blades 
there is a difference of pressure between the two sides of every 
blade and every row of blades; thus in Fig. 95, page 201, the 
low-pressure section is at the right-hand side of the illustration 
and consequently the pressure is greater on the left-hand side of 
every row of blades than on the right. It follows then that the 
rotor is subject to an axial thrust,* which must be balanced 
in order to make the unit operative. There is also a leakage 
of steam around the tips of the stationary and moving blades 
which must be minimized. The fact that the steam admission 
is over the complete circumference accounts for the reduction 
of the diameter of the drum toward the high-pressure end, in 

* In the example given, this thrust is exerted as a force pushing the rotor toward 
the right. 



STEAM TURBINE^TYPES AND BLADE DESIGN 113 

order to obtain blades of appreciable heights and also to reduce 
the steam leakage mentioned above (see Fig. ii2i). The factors 
governing leakage and its influence on design will be discussed 
in detail in succeeding pages. 

The effort to eliminate or reduce the difficulties enumerated 
above accounts for the various designs of reaction turbines now 
on the market. To eliminate axial thrust, the double flow ar- 
rangement shown in Fig. 184 and Fig. 112J has been developed 
for the larger sizes. To avoid small blade heights and to reduce 
leakage and axial unbalance in units of small or moderate size 
(500 to 12,000 kilowatts) accounts for the combined impulse and 
reaction type shown in Fig. ii2g. 

The diameter of the low-pressure section of the rotor of a 
reaction turbine is determined by the permissible blade speed 
and the rated speed of rotation (revolutions per minute) . With 
a drum construction it is not permissible to adopt peripheral 
speeds for the rotor higher than about 600 feet per second. The 
speed of the rotor (revolutions per minute) will depend on the 
capacity of the turbine, or more particularly, if it is to be con- 
nected to an electric generator, on the allowable speed of the 
generator. 

A table * on page 114 gives the rated speeds of a number of 
different sizes of commercial turbines of the Parsons type, some 
of which, it will be observed, are not for the standard frequencies 
of alternating current generators used in America. It is much 
more difficult to design an electric generator with as much 
strength in the rotating field or armature as in the turbine parts. 

The diameter of the low-pressure section of a simple reaction 
turbine is generally made V2 times that of the intermediate sec- 
tion, and the diameter of the intermediate section is V2 times 
that of the high-pressure section. It follows then that the ratio 
of the blade speeds of successive sections is also v 2.| 

The speed of rotation of turbines direct connected to alternat- 
ing-current electric generators is usually determined by the 

* Trans. Inst, of Engineers and Shipbuilders (1905-06). 
t In some recent American designs this ratio has been increased to more than 1.5. 
See designs on pages 138 to 141. 



U4 



THE STEAM TURBINE 



frequency of the alternating current. For the usual frequency 
in America for electric lighting (60 cycles per second) the gen- 
erator must be operated at 3600, 1800, or 900 revolutions per 
minute; and for 15 cycles per second the revolutions cannot, of 
course, exceed 900 per minute, as a generator cannot be built 
with less than two poles. 



Normal Output of Turbine. 



250 kilowatts 
500 kilowatts 
750 kilowatts 
1000 kilowatts 
1500 kilowatts 
2500 kilowatts 
3500 kilowatts 
5000 kilowatts 



Peripheral Blade Speed, 
Feet per Second. 


Number of 

Rows 

of Moving 

Blades. 


First 
Expansion 
(Section). 


Last 
Expansion 
(Section). 


IOO 
120 
125 
125 
125 
125 
138 
135 


2IO 
285 
260 
250 
360 
300 
280 
330 


72 
60 

77 
80 
72 
84 
75 
70 



Revolutions 

per 

Minute. 



3000 
3000 
2000 
1800 
1500 
1360 
I200 
750 



The English method of designing purely reaction turbines of 
the Parsons type will now be discussed. It will be observed 
that in this foreign engineering practice much lower blade speeds 
are used than in the designs just explained. 

Usually in these English designs with three different diameters 
of the rotor (three sections), the number of rows of blades or 
stages is arranged so that one-quarter of the total work is done 
in the high-pressure section. The intermediate section takes 
also one-quarter of the total work, and the low-pressure section 
one-half. 

Calculation of Number of Stages. The force F on a moving 
blade in the direction of motion is for W pounds of steam flow- 
ing per second, Fig. 55a * as follows: 



W 



(27) 



F = — (V 2 cos a + V 3 cos /3).f 
g 

* Fig. 55a is constructed on the assumption that the stationary and moving 
blades in a stage are to be similar in outline and angles. The triangles as shown 
at the entrance and exit sides of the moving blade are then equal with corresponding 
angles indicated. (See Morrow's Steam Turbine Design, page 114.) 

t The absolute rather than the relative velocities are chosen for this discussion 
because they are more adaptable Tor the simplification of algebraic equations. 



STEAM TURBINE TYPES AND BLADE DESIGN 115 

Work done by W pounds of steam per second is in foot- 
pounds per second per stage, 



— (V 2 COS a + V 8 COS |8)V 6 . 



(27a) 




a V 6 d c 

Fig. 55a. Diagram for Calculation of Number of Stages. 



Work done per pound of steam in foot-pounds per second per 
stage is, then, 

1 



g 



(V2Cosa + V 3 cos/3) V&. 



(27b) 



The following relation is also easily established 
V& = V 2 cos a — V 3 cos jS. 

Work per pound of steam, foot-pounds per second per stage, 
can then be written in the form 



-(V 2 2 cos 2 cx- F 3 cos 2 0). 
g 

The following trigonometric relations are obvious: 

. _ V 2 sin a . 2 _ V 2 2 sin 2 a , 

sin j8 = — — — ; sin 2 /? = — — — ; and 



(27c) 



cos 2 /3 = 1 — 



V 2 2 sin 2 a 

v 3 2 



Ii6 THE STEAM TURBINE 

Substituting this last value in equation (27c) we have 
Work per pound of steam (foot-pounds per second) 

= - ( V 2 2 cos 2 a - Vi + V 2 2 sin 2 a) 

g 

= -[F 2 2 (cos 2 Q ! + sin 2 a) - F 3 2 ] 
g 

= - (V 2 2 - V3 2 ). (2 7 d) 

g 

The last equation might have been written directly in kinetic 
energy relations, without the preceding steps. 

Using the right-angled triangle (bed), of which V 3 is the hypot- 
enuse, 

Vi = V 2 2 sin 2 a + (V 2 cos a - V h ) 2 

Vi = V 2 2 sin 2 a + V 2 2 cos 2 a - 2 V 2 V b COS a + V b 2 

V 3 2 = V 2 2 - 2 V 2 V 6 cos a + V 6 2 . (27e) 

Substituting (27c) in (27d) there results, 

Work per pound of steam per stage (foot-pounds per second) 

= L(V 2 2 -V 2 2 + 2 F^coso: - V b 2 ) 
g 

= - (2V 2 COSa- V 6 ). (27f) 

g 

The ratio — 7 is often represented by k, then 

Work per pound of steam per stage (foot-pounds per second) 

V& 2 
= — (2 k cos a — 1). (27g) 

g 

This value can be equated to the mechanical equivalent of the 
total heat available E a or 

V* 2 , v E a X 778 

(2 k COS a — i) = ' 7 

g N 

where N is the number of stages in the section or the turbine, 



STEAM TURBINE TYPES AND BLADE DESIGN 117 

as the case may be, in which the available energy E a is utilized. 
Solving for N, we have 

N * = E„X778Xg ( . 

If the heat drop and the average value of k are assumed to be 
the same in each expansion or stage, and blade angles are also 
approximately constant, then 

NW = a constant. (271) 

Example. Make approximate calculation of the number of 
stages (N) for the high-pressure section of a steam turbine in 
which the heat drop in the section is 200 B.T.U. per pound of 
steam, assuming k = 2.5, cos a = .94, V 6 = 200, and efficiency 
of energy transformation is 70 per cent. 

Using equation (27I1), corrected for efficiency, we obtain 

N = -70 x 200 x 778 X 32.2 = 

(2oo) 2 (2 X 2.5 X .94 - 1) 

Designers of marine turbines usually assume the value of the 
constant in equation (271) above at about 1,500,000 to 1,600,000; 
but for electric generator service, where much higher peripheral 
speeds are allowable, the value of this constant varies from 
2,200,000 to 2,600,000, depending somewhat on the allowable 
radial clearances. The lower value can be used when the ma- 
chine work is accurate and the designing has been done with 
great care to eliminate unequal expansion between the rotor and 
the casing (see page 132). 

Gauging Blades. It sometimes happens when arranging the 
blading in groups, that a fractional part of a stage is shown by 
the calculations. In such a case two groups may be combined 
into one of about the average height, if in this way a whole 

* Observe that no losses have been allowed for in the discussion of the equa- 
tions leading to this result. In actual calculations of designs a coefficient of effi- 
ciency, usually .70, is inserted in the numerator as in the example following. 
The expression NV& 2 has useful applications in coordinating results of tests on a 
complete line of machines and in obtaining good approximations of efficiencies for 
new designs. 



Il8 THE STEAM TURBINE 

number of rows can be secured. Probably it will then be found 
that one or two of the last rows of blades do not give sufficient 
area for the passage of the steam, and this area is then increased 
by " gauging " the blades in both the rotor and casing. This 
" gauging " is done by forcing a piece of metal — preferably not 
much harder than the metal of the blades — between the blades 
so as to twist them more nearly parallel to the axis. Manufac- 
turers using steel blades have usually special keys made for the 
purpose of twisting the blades by hand both for the purpose ot 
" gauging " and for changing the blade angles in order to secure 
an accurate balance between the end thrust of the balance pis- 
tons (see page 198) and that of the blades. It is stated on very 
good authority that this twisting of the blades and changing 
the angles for the purpose of balancing end thrust as much as 
5 degrees does not appreciably alter the economy of the turbine. 

In the last rows of the low-pressure section the discharge or 
outlet angle (7 in Fig. 47b) is invariably increased beyond the 
calculated value to permit the flow of a large amount of steam 
through the blades without choking and still maintain a constant 
blade height in these rows. It has been shown that the area of 
the passage for steam at the discharge side of blades is wD sin 7 
X k f * if blade thickness is neglected. Hence, with constant 
values of D and h, an increase in the value of the discharge angle 
7 is an increase in the area for the flow of steam. It is thus 
easily possible to avoid inconveniently long blades in the low-pres- 
sure section by increasing 7 and keeping the blade height constant. 

It will now be shown how the angles required to maintain 
constant blade height can be calculated if the actual specific 
volume of the steam in the various rows of blading can be deter- 
mined. The volume of steam flowing in cubic feet per second is 
tD sin 7 X h X F^.f And if v is the specific volume of the 
steam and W is the flow through the blades in pounds per second, 
we have y^^ sin y = Wv# (ag) 

* See page 98. This is for steam flow over the whole periphery as is customary 
in reaction turbines. 

t Vr2 is the " relative " velocity in the blades as used in Fig. 47b. 



STEAM TURBINE TYPES AND BLADE DESIGN 119 

For the expansions (stages) in which W, D, and h are to be 
constant, the equation simplifies to 

Vr2 sin 7 proportional to v. (28a) 

It is desirable in such cases to have constant values for the 
component of the velocity ( V r2 cos 7) * of the approaching steam 
in the direction of rotation of the blades; that is, in the direc- 
tion of Vb. In other words F r2 = a constant -5- cos 7. Making 
this substitution, we have 

= tan 7 proportional to v. (28b) 

cos 7 

When the average value of v is known, in the stage preceding 
those to be gauged, equation (28b) is used to calculate values of 
7 for the corresponding values of v of the steam in the gauged 
blades. 

The following example illustrates the method. Assume the 
specific volumes of the steam in the last three stages of a reac- 
tion turbine are respectively 90.5, 143, and 230 cubic feet per 
pound, and that the blade angle for the first row of this group is 
normal (without gauging) and that 7 = 20 . 

Now represent the discharge angles to be calculated by y 
and 7" '. Then tan y -f- tan 20 = 143 -r- 90.5 or y = 30 
(nearly). Also tan y" -f- tan 20 = 230 -J- 90.5 or y" = 43 
(nearly) . It has been thus determined that the discharge angles 
for the last two rows are to be 30 and 43 . Blades which have 
been adjusted in this way to increase, above the normal, the 
discharge angles are called wing blades. 

Considerations in Designing. An example illustrating the de- 
sign of a commercial type of reaction turbine will now be dis- 
cussed. 

The difficult part and that requiring the best judgment in the 
designing of a reaction type of steam turbine is in determining 
as accurately as possible the volume of steam that will pass 
through the blades for its full capacity; that is, when all the 

* Morrow's Steam Turbine Design, page 119. Vn cos 7 is sometimes called 
the " velocity of whirl " at the entrance to the blades. 



120 



THE STEAM TURBINE 



valves controlling the admission of steam are wide open; or in 
other words when there is no throttling of the steam pressure. 
It is for this flow that all the blades must be proportioned for 
their best efficiency. It is presumed that for both lighter and 
heavier loads the efficiency and the steam consumption will not 
be so good. In order to determine the volume of steam flowing 
at this condition obviously the actual number of pounds of 
steam to be used by the turbine must first be known. 

As the result of a great deal of study the curve shown in 
Fig* 55t> has been developed from data collected in a large part 













»" .' ^h*£^' 


$60 <fT ' 


33 .<£_ 




& ? 


*40 


'S 




gj.30 















50,000 



100,000 150,000 

Coefficient "C" 



200,000 



Fig. 55b. Efficiency Ratios for Reaction Turbines. 

by Martin * as applying to a large variety of turbines, but par- 
ticularly to the reaction type. This curve shows by its ordinates 
the so-called " efficiency ratio, " which is the ratio of the theo- 
retical steam consumption (see page 31) to the actual steam 
consumption (water rate) per kilowatt-hour. The values on this 
curve for values of the coefficient " C " f from 110,000 to 120,000 
are for reaction turbines either above 5000 kilowatts' capacity, 
or else for sizes between 1500 to 5000 kilowatts which have clear- 
ances at the tips of the blades too small for standard prac- 

* Design and Construction of Steam Turbines, 1913. 

f Values of this coefficient should not be confused with the " constant " given 
on page 117. 



STEAM TURBINE TYPES AND BLADE DESIGN 121 

tice. Many of the turbines of the latter or small sizes did, in 
fact, strip their blades a short time after being put into service. 
Values less than 100,000 are for sizes . smaller than 1500 kilo- 
watts. Other things being equal the smaller size turbine should 
have a lower value of the coefficient. Thus for a turbine of 
1000 kilowatts' capacity the proper coefficient should be between 
80,000 and 90,000 and for a 2000 kilowatt size the coefficient 
should be about 100,000. In fact the latter value is generally 
used by careful designers of Parsons types for all sizes from 
1500 to 5000 kilowatts, having reasonably large clearances at 
the tips of the blades. 

Practical Example. A reaction turbine with a drum rotor of 
three sections is to be designed to give a rated output of 2000 
kilowatts, operating at 1500 r.p.m. When supplied with steam 
at 165 pounds absolute pressure, ioo° F. superheat, and 1 pound 
absolute exhaust pressure (about 28 inches vacuum), the turbine 
shall carry 10 per cent, overload before the by-pass or overload 
valve (see page 307) opens. 

For this design (2000 kilowatts), therefore, the value of the 
coefficient should be 100,000. Ordinates of the curve in Fig. 55b 
show that the corresponding efficiency ratio is about .675. 

The available energy from the entropy-total heat chart from 
the initial conditions of 165 pounds per square inch absolute 
pressure and ioo° F. superheat to the final pressure of one pound 
absolute is 1252 — 908 or 344 B.T.U. per pound of steam. 
Dividing the B.T.U. equivalent of a kilowatt-hour, which is 
3412, by the available energy per pound of steam (344) we 
obtain a theoretical steam consumption of 9.92 pounds. This 
theoretical steam consumption divided by the efficiency ratio 
gives the actual steam consumption of the turbine per kilowatt- 
hour as measured at the turbine shaft or 9.92 -f- .675 is 14.69 
pounds. If we assume 5 per cent, loss for generator and con- 
nections to the switchboard then the steam consumption per 
kilowatt " at the switchboard " is 14.69 -5- .95 or 15.46 pounds 
of steam when dry saturated. 

If the steam is superheated, as in this case, the steam con- 



122 THE STEAM TURBINE 

sumption of reaction turbines is reduced, further, at least 6 per 
cent, for ioo° F.* (see page 169), so that the actual number of 
pounds of steam to be passed through the turbine for the condi- 
tions stated for this design is 15.46 X .94 or 14.54 pounds per 
kilowatt per hour " at the switchboard." 

The turbine must be designed for a total steam consumption of 
14.54 X 2000 X 1. 10 = 31,993 pounds per hour or 8.88 pounds 
per second at maximum output (10 per cent, overload), when the 
admission valve will be wide open so that there is no throttling. 
Then the steam entering the first row of blades will be at 165 
pounds absolute pressure, of which the volume at ioo° F. super- 
heat is 3.21 f cubic feet per pound. The volume of steam ad- 
mitted to the turbine per second is 3.21 X 8.88 = 28.50 cubic 
feet, and just as in the design of impulse turbines, the blades 
must be designed for the passage of this amount of steam. 

The blades are designed by determining the entrance and dis- 
charge angles by velocity triangles like those in Fig. 46 after 
the available energy for each stage has been calculated. Some 
designers make their calculations for the rated full load condi- 
tions and not for the maximum output obtained just before the 
stage valve opens. The difference between the two methods is 
that until the maximum output is reached, without opening the 
stage valve, there is obviously some throttling in the admission 
valve,{ and when designing for full load conditions this throttling 
must be allowed for. For this reason it is preferable to design 

* Two allowances for superheat are thus made. The steam consumption as 
calculated above was determined by using the theoretical available energy including 
the superheat. The correction which is now made is to compensate for the effect 
of superheating in reducing rotation losses. 

t Marks and Davis' Steam Tables and Diagrams. In these tables the specific 
volumes have been calculated by Knoblauch's equation, which gives considerably 
larger values than equation (9) . The results of different investigations do not give 
any sort of agreement, the rate of increase of volume with superheating varying 
as much as 100 per cent. It is usually stated that the specific volume of super- 
heated steam is 15 per cent, larger for ioo° F. of superheat than that of dry sat- 
urated steam. According to Knoblauch's equation used by Peabody, this per- 
centage is about 17, and according to equation (9) it is about 13. 

% There is some throttling even in the " blast " or " pulsating " valves (page 
295) used in nearly all types of Parsons turbines. 



STEAM TURBINE TYPES AND BLADE DESIGN 



123 



for maximum output when the admission valve must be wide 
open.* The available energy is then calculated by steps from 
the rated admission to the exhaust pressure. This available 
energy might be determined for every stage as it is done for 
designing impulse turbines, but this is unnecessarily laborious, 
as the pressure drop is so small. Approximately the same result 
is obtained by calculating assumed expansions, in stages of 10 
B.T.U. with the same reheating factors as would be used for the 
same size of impulse turbine. For a 2000 to 3000 kilowatt size 
the reheating (page 102) should be not much more than 30 per 
cent. Assuming this value, the total available energy as read 
from the entropy-heat chart with reheating for every 10 B.T.U. 
from 165 pounds absolute and ioo° F. superheat to 1 pound 
absolute exhaust is 260 B.T.U. f Without considering reheating 
it would have been 343 B.T.U.; but with 30 per cent reheating 
as if in only one "step" it is only 240 B.T.U. The quality of the 
steam in the last stage after reheating "by steps" of 10 B.T.U. 
is .886. 

For this design it will be assumed that in using equation (271) 
on page 117, 

NV b 2 = 2,560,000. 

It has already been stated that the diameters of the sections of 
the rotor of this type of reaction turbine increase as V2 ; and as 

* The steam consumption at full and fractional loads can be estimated by- 
drawing a " Willans" line of total steam per hour (page 160). Unless the design 
of a steam turbine is radically wrong, usually because of insufficient area of the 
steam passages, which is called " choking" the steam, it has been shown by expe- 
rience that the points representing total steam per hour plotted against fractional 
loads will be on a straight line from no load to the maximum output (without a 
stage valve). At no load a Parsons turbine usually takes one-eighth of the total 
quantity required at the normal maximum output. By plotting these two points 
(no load and maximum output) and joining them with a straight line, the total 
steam consumption at all other loads can be read and the steam per kilowatt-hour 
or per horsepower-hour can be calculated with considerable accuracy. 

t This available energy should be read in the same way as for the design of the 
impulse turbine illustrated in Fig. 52; meaning, that the energy should be obtained 
by subtracting from the total heat at the initial condition of pressure and superheat, 
the total heat at the final pressure, without the last reheating. There are some 
designers of impulse turbines, however, who use the calculated net available energy 
after reheating in each stage; but the problem then becomes very complicated, as 
most of the reheating takes place after the steam is discharged from the nozzles 
or stationary blades. 



124 



THE STEAM TURBINE 



the blade speeds must increase in the same proportion as the 
diameters, the following speeds of the blades will be assumed, 
which are not at variance with good practice : 

Vb of first section of rotor = 140 feet per second. 

Vb of second section of rotor = 200 feet per second. 

Vb of third section of rotor = 280 feet per second. 
The value of peripheral speed, 140 feet per second for the first 
section of the rotor, corresponds at the speed of rotation required 
(1500 revolutions per minute) to a diameter of 

— ^— =1.78 feet or 21.36 inches. 

3.1410 X 1500 

It is stated by Martin that English designers of reaction turbines 
for land service between 1000 and 6000 kilowatts' capacity deter- 
mine the diameter of the first section of the drum of the rotor di 
by the following empirical formula based on experience: 

-3 410,000 WVp 

r.p.m. 

where w is the weight of steam flowing through the turbine in 
pounds per second at maximum output without the stage valve 
being open, and v is the specific volume of the steam at the con- 
dition it enters the turbine in cubic feet per pound. In this case 
the diameter of the first section of the rotor as calculated by this 
formula would be 

, , 410,000 X 8.88 X 3.21 
> _^° I5 oo = 779 °' 

and di is v 7790 or 19.8 inches, which agrees well with the value 
calculated (21.36 inches) from the assumption of a satisfactory 
peripheral speed. Actually it is better practice and certainly 
more rational to assume a safe peripheral speed than to deter- 
mine the diameter by formulas having important empirical 
coefficients which must vary necessarily considerably with the 
type of the design. Allowable peripheral or blade speeds are 
always about the same for a given speed and type of construc- 
tion. Limits as regards peripheral speeds can be as accurately 
determined as any other problem in the designing of machines. 
A similarly empirical formula is sometimes used by designers 
of reaction turbines to determine the least permissible diameter 



STEAM TURBINE TYPES AND BLADE DESIGN 125 

of the low-pressure section in the last stage, with the object of 
reducing to a minimum the losses due to excessive residual 
velocity of the steam as it discharges into the exhaust pipe. If 
the diameter of the rotor measured to the middle of the blades 
in the last stage is d z then for this reason (d 2 ) 2 should be not less 
than .57 X the output when there is no throttling in the main 
inlet valve. In this case then 

d* 2 > -57 X 2000 X 1. 10 (see page 122), 
or d 2 must be at least 32.07 inches, which corresponds to a periph- 
eral speed of ^ ^ — , or 210 feet per second. 

F 12 X 60 ^ 

The value selected for this stage (280 feet per second) from the 
viewpoint of permissible stresses is well above this minimum 
limit. Blade speeds as high as 600 feet per second are now used 
in some American designs of steam reaction turbines in large sizes. 

Having determined that the conventional blade speeds are 
very satisfactory for this design the required number of reaction 
stages will be calculated in the usual manner as follows: 

If the blade speed of the whole turbine had a constant value of 
140 feet per second, then 

(140) 2 X N = 2,560,000; N = 128 (nearly). 
As, however, only one-fourth of the work is to be done by the 
first section * operating at this blade speed, the number of stages 

t oR 

in the first section is = 32. The value of N for the second 

4 
section is 64; and as one-fourth of the work is done also in this 
section, the number of stages is 16. For the third section N is 
32, and since one-half of the work is done in this section, the 
number of stages is 16. 

Each section of the rotor is commonly divided into two or 
four groups or " expansions." 

Reaction turbines are usually designed for equal work (energy) 
per stage for a given section of the rotor. In the high-pressure 
or first section, one-quarter of the work is done, and the avail- 
able energy for each of its thirty- two stages is 260 B.T.U. •*■ 
(4 X 32) = 2.03 B.T.U. Similarly the available energy for 

* See page 114. 



126 THE STEAM TURBINE 

each stage of the intermediate section is 260 B.T.U. -5- (4 X 16) 
= 4.06 B.T.U. ; and for each stage of the low-pressure section 
is 260 B.T.U. ^ (2 X 16) = 8.13 B.T.U. It may be assumed 
that about one-half of the available energy in each stage produces 
velocity in the stationary blades and the other half in the moving 
blades. The theoretical angles are determined from velocity tri- 
angles, applying the coefficients from curve B in Fig. 51, by the 
usual methods as explained for impulse turbines. The discharge 
angles for all the stages except the last groups in the low-pressure 
section will be assumed to be 20 degrees. The angles for the last 
stages will be made 45 degrees. It is obvious, of course, that 
the discharge angle is always the same as the " absolute " angle at 
which the steam enters the succeeding row of blades. In this 
design no allowances are made for probable " gauging " of the 
blades to adjust the thrust on the rotor or for other reasons.* 
The velocity of the steam leaving the first row of stationary 
blades in the high-pressure section is about 225f feet per second. 
A net area of 28.50 cubic feet -J- 225, or .127 square feet, or 18.2 
square inches, is required to pass the steam. As the discharge 
angles of the blades in the high-pressure and intermediate sec- 
tions are to be made 20 degrees, that value will be taken for this 
design, and the actual area of the blade ring will be approximately 
18.2 -7- .342,} or 53.3 square inches. 

The blade speed of the high-pressure rotor is 140 feet per 
second,§ so that the mean diameter of the blade ring is 

* It has been stated that some makers of marine turbines who have not had 
much experience in building them vvill often design turbines to give considerably- 
larger output than is intended for the service and then reduce the output to the 
required rating by " gauging " the blades. 

t In a reaction turbine the maximum velocity in each stage is attained when the 
steam is discharged from the stationary blades. Although there is expansion also 
in the moving blades, more velocity is absorbed in them than is produced, and the 
velocity of the steam discharged from the moving blades is considerably less than 
225 feet per second. 

% The total area of the annulus for blades with discharge angles of 20 degrees is 
the net required area divided by sin 20 degrees (see Fig. 49). Practical designers 
often call the sin of 20 degrees one-third and make the area of the annulus three 
times the net required area. 

§ Manufacturers generally appreciate the gain from operating at high peripheral 
speeds of the rotor. To-day efforts are directed generally by all makers of direct- 
connected turbine-generators to improve the mechanical construction of the gen- 
erator to run at higher speeds. 



STEAM TURBINE TYPES AND BLADE DESIGN 127 
140 X 60 



, = 1.78 feet, or 21.4 inches, and the height of the 
1500 X 3.1410 

first row of blades on the rotor is approximately 53.3 square inches 
-s- 21.4 X 3. 141 6 = .80 or nearly yf inch (see table, page 128). 
With full rated pressure in the admission chamber about 7 per 
cent, of the total steam leaks through the " dummies " or balance 
pistons * at the high-pressure end of the turbine. This leakage 
as well as that around the tips of the blades through the radial 
clearance is not considered here in the calculations. It is prob- 
able, however, that the amount of this leakage is quite sufficient 
to allow for the thickness of the blades on the discharge side.f 
The volume of the exhaust steam (1 pound per square inch abso- 
lute pressure and .886 quality) is 297 cubic feet per pound. 
Initially the volume was 3.215 cubic feet per pound, so that the 
volume in the last row of blades is 92.5 times that at admission. 
Since one-fourth of the work is done in the blades of the first 
section, one-fourth of the total expansion occurs in them, or the 
volume entering the second section is Vg 2 .5,{ or 3.10 times the 
original volume. Since the mean diameter is to be made V2 
times that at the high-pressure end and the steam velocity is to 
be also V2 times as great so as to correspond with the increase 
in blade speed (which is V2 times that in the first section. See 
page 113), the height of the blades in the first row of the inter- 

mediate section will be °' — =1.55 times that of the first 

V 2 X v 2 

row in the high-pressure section. Similarly the blade height for 

the first row of the low-pressure end will be 1.55 times that of the 

first row of the intermediate section. Each of these sections will 

be divided into four groups or " expansions." Since the volume 

is increased four times for each section, the blade height of each 

* Methods of calculating the leakage through the balance pistons and that through 
the radial clearance of the blades in reaction types are discussed on pages 135-139. 

t Thomas uses a coefficient of 1.5 to increase the area of the blades to allow for 
the thickness at the discharge side. If the blades are made thin at their edge^, as 
in good designing, it is not customary to use a coefficient " for the thickness of the 
blades." 

t Let v' = volume at end of third section, 

v\ = volume at beginning of first section 
x = number of expansions, 
then v' = vi x , and Vi = *\fv'. 



128 



THE STEAM TURBINE 



of the high-pressure and intermediate groups will be ^3.10, or 
1.33 times as large as in the preceding one. 

The calculated blade heights for each of the four groups of the 
high-pressure and intermediate sections are given in the follow- 
ing table: 





Group Number. 




1 


2 


3 


4 


Blade height, high-pressure section 
Blade height, intermediate section 


if 


is 


2h 


iM 
2H 



Blade heights are adjusted to sixteenths, although in practice 
the nearest eighth is commonly used. 

Because of the long blades in the low-pressure section they will 
be made in eight groups. The height of the first group will be 
1.55 times the height of the first group of the intermediate section. 

The volume entering the third section is ^92.5 X ^92.5 = 
9.61 times the original volume, and blade height in first row of 
third section is 1.55 X 1.55 X height of first row = (1.55) 2 X .80 
= 1.9 inches, or approximately ill inches. 

Each blade in third section is ^9.61 = 1.33 times height of 
preceding one, or the height of second row is 1.33 X ijf or 2.57 
(approximately 2 A inches) . The results are tabulated as follows : 





Group Number. 




1 


2 


3 


4 


5 


6 


7 


8 


Blade height (inches) 


iH 


2& 


3T6 


A 9 
4l6 


6^ 


8| 


iof 


I4l 



1 Martin states that at the high-pressure end (in turbines for 
stationary service) it is desirable to limit the blade height to 
not less than one-twenty-fifth (-fa) of the drum diameter. If 
the blades are shorter than this the loss by leakage around the 
tips may become excessive. In marine turbines the high-pressure 
blades in the first section are only tV of the drum diameter. 
On this basis the blade heights might be slightly increased. 



STEAM TURBINE TYPES AND BLADE DESIGN 



129 



At the low pressure end of the turbine the length of the blades 
would be considered excessive in practice. It is a rule generally 
followed by designers of reaction turbines not to make the 
greatest blade height more than one-sixth the mean diameter of 
the blades for the section considered. The mean diameter of 
the low pressure section is 21.4 X V2 X V^2 = 42.8 inches. 
In this case the maximum height would be, therefore, about 7.1 
inches. In order to reduce the length of the blades so that 




Fig. 56. Details of the Design of Reaction Blades. 

practical requirements shall not be exceeded, the discharge angle 
of the blades must be made greater than 20 degrees. Such blades 
with enlarged "exit" angles are called wing blades. The tangent 
to the curve at the back of the blade on the entrance side becomes 
about 90 degrees, and at the discharge side 45 degrees instead 
of 20 degrees. As the result of this change, the net area for the 
passage of the steam is .71 * (sin 45 degrees) instead of the 
standard "J"f of the annulus without blades. Wing blades 
7 inches long can be used to replace satisfactorily the blades in 
the 5th group; but as those of the 6th, 7th, and 8th groups must 

* In the turbines of the steamer Mauretania, wing blades giving a passageway 
of .86 of the annulus were used, but such a large degree of " winging " is not 
adopted in steam turbines for electric generators. 

t The sin of 20 degrees is .34, but practical designers take it often for convenience 
in calculating as $. 



i3° 



THE STEAM TURBINE 



be made of the same length, these blades will be shorter than they 
should be. This constriction of the steam passage, however, can- 
not well be avoided without making the rotor in four diameters. 
Fig. 56 shows how the blades of reaction turbines are laid out. 
As explanatory of this figure a table is given below showing the 
corresponding dimensions used by one manufacturer.* In the 
table data for five standard blades are given for varying discharge 
angles (0) from 20 degrees to 35 degrees and blade widths (w) 
of .25, .375, and .50 inch. All the linear dimensions are given 
in inches. /? is the entrance angle of the blades. 



Blade Number (Arbitrary). 





■ 


2 


3 


4 


s 





20 


20 


20 


30 


35 
1 8° 40' 


a 


10° 


9° 30' 


M° 30' 


1 5° 45' 


w 


0.25 


o-375 


0.50 


0.50 


0.50 


ft 


67° 30' 


67 30' 


67° 30' 


6o° 


6o° 


R 


0.485 


o.555 


0-794 


0.804 


0.810 


A 


o-035 


0.045 


O.068 


0.050 


0.040 


b 


0.020 


0.020 


0.020 


0.020 


0.020 


*, 


0.172 


0.260 


O.342 


0.313 


0.304 


C 


0.049 


0.040 


O.IIO 


0.147 


0.218 


R 2 


0.070 


0.109 


0.164 


0.210 


0.212 


k 


0.008 


0.010 


0.015 


0.040 


0.056 


m 


0.123 


0.185 


0.288 


0.280 


0.280 


n 


0.185 


0.280 


0383 


0.332 


0.330 


I 


0.166 


0.223 


0.282 


0.156 


0.134 


H 


0.478 


0-55 2 


0.770 


0.770^ 


0.770 



Another table is given here showing the principal dimensions 
of a 400-kilowatt reaction turbine with 3, 4, and 5 groups per 
section. The blade numbers in this table refer to the corre- 
sponding numbers in the table above. This table is partic- 
ularly useful for showing values assumed by designers for the 
blade pitch. It is not considered practicable in this type of blade 
construction to use a smaller pitch than .177 inch when a calking 
tool must be inserted between the blades. Manufacturers have 
usually curve sheets of empirical data from which the pitch and 
other standard dimensions are obtained. 

* The Engineer, Dec. 16, 1907. 



STEAM TURBINE TYPES AND BLADE DESIGN 131 
FIRST SECTION. 



Number 


Diameter of 


Dis- 


Blade 


Blade 
Number 


Volume 




Blade 
Pitch. 


Number 


of 


Section in 


charge 


Height in 


Cubic Feet 


V h + K 2 


of 


Group. 


Feet. 


Angle. 


Inches. 


per Pound. 




Blades. 


1 


0.84 


20 


O.6875 


■ 


4.08 
5-6 


.62 


O.I77 
O.25 


179 

J2 7 


2 


O.84 


20 


I. OO 


I 


5-6 

7.38 


.62 


0.1875 
0.2475 


169 
128 


3 


O.84 


20 


I.25 


1 


7.38 
8.92 


.62 


O.172 
O.2175 


184 
146 



SECOND SECTION. 



I 


1. 187 


20 


0.6875 


2 


8.92 
10.63 


.62 


0.20 
0.31 


180 
116 


2 


1. 187 


20 


o.9375 


2 


10.63 
13.8 


.62 


0.215 
0.3075 


207 
144 


3 


1. 187 


20 


1.2 


2 


13* 
18.8 


.62 


0.215 
o.3 2 3 


207 
i S 8 


4 


1. 187 


20 


i-75 


2 


18.8 
26.6 


.62 


0.208 
0.326 


214 
J 37 



THIRD SECTION. 



I 


1.88 


20 


o.9375 


2 


26.6 
35 -o- 


.62 


0.208 
0.307 


34o 
230 


2 


1.88 


20 


i;3i25 


2 


35 -f 

52.8 


.62 


0.208 
o.339 


340 
210 


3 


1.88 


20 


2.00 


2 


52.8 
83.8 


.62 


0.198 
o.355 


358 

200 


4 


1.88 


30 


2-75 


4 


83.8 
161 


.69 

•55 


0.25 
0-53 


284 
i34 


5 


T.88 


3o 


4-5 


4 


161 

3 11 


.70 
.46 


0.308 
0.54 


230 
J3 1 



Radial Leakage. As the volume of the steam increases, the 
area of the annulus of each ring of blades must, of course, increase 
proportionally. This increased area would be obtained most 
easily, as with impulse turbines, by increasing the blade heights 
in each ring. This method, however, would make it necessary 
to carry as stock in the store-room a great number of blades of 
different sizes. In order to reduce the stock of blades and to 
reduce the cost of machining the rotor and casing, it is custom- 
ary to make a group of several rows of blades of the same height, 



132 



THE STEAM TURBINE 



and the required increase in area through each ring of blades is 
obtained by decreasing the number of blades in each succeeding 
stage. The two values of volume, pitch, and number of blades 
given for each group in the preceding table are for the rows at the 
beginning and at the end of the group. 

In the discussion of the design of these reaction turbines it has 
been assumed that each section of the rotor is made of the same 
diameter from the first to the last group. For theoretical con- 
siderations this assumption is permissible, but actually for each 
blade group the diameters of both the rotor and casing are 
changed so that approximately half the increase in blade height 
is cut out of the casing and the other half is taken from the rotor. 
It is usually stated that this is done merely for mechanical reasons, 
but this method has advantages also in order to secure the best 
steam flow. It is well known that steam tends to fill completely 
the passage through which it flows and at the same time expand 
at right angles to its axis of flow. Now if all the expansion is 
made on the casing side of the blades the expansion of the steam 
will increase the leakage around the tips of the blades next to the 
rotor without materially affecting the leakage at the tips nearest 
the casing. 

The leakage of steam around the tips of the blades depends, 
of course, again upon the amount of the radial clearance. Im- 
provement in the design of reaction turbines will be largely 
accomplished (i) by skillful designing and machine work to 
permit the reduction of radial clearances and (2) by increasing 
the blade speed. In fact the question of allowable radial clear- 
ance depends finally on the blade speed. If the blade speed is 
increased it is possible to use higher steam velocities with larger 
pressure drop per stage, and consequently fewer stages. This 
is apparent also from an inspection of the designing formula on 
page 117. With fewer stages a shorter rotor is required which 
will also be proportionately stiff er; and with a stiff shaft it is 
possible to allow very small radial clearances, provided, of course, 
temperature effects are carefully studied. 

The reader will have observed that the design of reaction 



STEAM TURBINE TYPES AND BLADE DESIGN 



133 



turbines is largely by " cut and try " methods. For this reason 
it is a financial absurdity for a manufacturer to-day to begin 
making reaction turbines unless he has practically unlimited 
resources and can obtain from makers of similar machines at 
not too large a cost the results of their experiences. 

The method explained here of determining the important and 
unique parts in the design of a reaction turbine for a given set 
of conditions, as regards maximum output, steam consumption, 
pressure, superheat and vacuum, although very simple in all 
essentials as regards standard practice, gives results on which it 
is impossible to improve by the most elaborate mathematical 
analysis imaginable. In fact all elaborately mathematical anal- 
yses of the action of steam in a reaction turbine depend finally 
on the substitution of certain coefficients, most of which have 
no basis in fact. 

The work done per stage is always much greater in current 
practice in impulse than in reaction turbines. For the same 
total limits of pressure the work per stage is inversely propor- 
tional to the number of stages. 

In general, we may say that mechanical considerations and 
the speed at which machinery can be conveniently operated de- 
termine the size and number of revolutions at which a turbine 
can be run. In a good design about the same total efficiency 
is obtained, whether the turbine is classified as an impulse or a 
reaction machine. 

NOTES ON THE DESIGN OF BLADING FOR COMBINED IMPULSE 
AND REACTION TURBINES. 

The rotor of a modern reaction turbine of moderate size con- 
sists of an impulse wheel with two moving rows of blades (velocity 
stages) followed by an intermediate and a low-pressure section 
carrying reaction blades. The mean diameter of the blades in 
the intermediate section is of course appreciably smaller than 
in the low-pressure section. The low-pressure sections are 
single or double flow, depending on capacity. The peripheral 
velocity of the impulse wheel varies between 400 to 500 feet per 



134 THE STEAM TURBINE 

second or over, depending on the available energy utilized in 
the first expansion. The higher velocities are used in the 
smaller sizes in order to reduce the steam pressure to a value 
suitable for the following reaction blading. The blade velocity 
at the mean diameter of the Parsons blading in the intermediate 
section is about 350 feet per second and in the low-pressure 
section 500 to 600 feet per second, depending on size and revolu- 
tions per minute. 

The purely reaction turbine for high-pressure steam is often 
preferred in America in very large sizes, 25,000 kilowatts and over, 
in which case the machine may be constructed in two separate 
casings. There are then really two distinct rotors (see Fig. 
112J), the high-pressure rotor having its blading in two sec- 
tions of different diameters and the low-pressure rotor, generally 
double flow, having its blading in one or two sections. The 
blade velocity in the high-pressure rotor varies between 250 to 
300 feet per second and in the low-pressure rotor between 300 to 
500 feet per second. 

Purely reaction turbines are also ideally suitable for either non- 
condensing or low-pressure operation (see Fig. 11 21). Very 
efficient units can be built having reaction blading only, for 
non-condensing operation at 3600 revolutions per minute or 
over and in capacities of 1000 to 2000 kilowatts or over. The 
available energy per stage (as explained in previous chapters) is 
proportional to the square of the peripheral velocity of the 
moving blades, and formulas for determining the available en- 
ergy per stage have been given. Having settled on the blade 
velocities in the various sections, the number of stages is deter- 
mined to utilize the complete range of available energy. The 
blade height in each section should of course increase gradually 
toward the low-pressure end. 

In practice, however, the same blade height is maintained 
for a group of 4 to 6 or more stages, as it is found that the 
necessary increase in area to allow for steam expansion can be 
obtained by gauging ; that is, by changing slightly the discharge 
angle of the stationary and moving blades. 



STEAM TURBINE TYPES AND BLADE DESIGN 135 

The discharge angle of reaction blading is normally about 20 , 
except in the last row of a condensing unit the discharge angle 
may be 35 to 40 . The edge of the blade on the discharge side 
is usually extended to give a definite discharge passage between 
blades. The entrance angle is not far from 90°, say 70 . The 
minimum blade height is about 1 inch and is increased by 
steps of J inch or 1 inch up to 16 inches or more. The limiting 
height is determined by allowable stresses which are readily 
determinable. The blade width varies from f inch in the lower 
heights and gradually increases to ij inches or i| inches, de- 
pending on the height. The pitch of the blades is about .8 of 
the width. The radial clearance is, of course, low in order to 
reduce leakage ; it varies between -£% inch and y 1 ^ inch or slightly 
more, depending on the height of the blades. Axial clearances be- 
tween stationary and moving rows are quite large, sometimes 
half the width of blades. 

FACTORS GOVERNING LEAKAGE AROUND STATIONARY AND 
MOVING BLADES. 

As stated previously there is a pressure drop in each row of 
stationary and moving blades; therefore a certain amount of 
steam flows axially through the clearance spaces between the 
tips of the moving blades and casing on one side and tips of 
stationary blades and rotor on the other side. If x is the radial 
clearance and D the diameter at the tip of the blades the leakage 
area is equal to Dx. The weight W of leakage steam flowing be- 
tween rows can be estimated by means of the following funda- 
mental relation which forms the basis of all nozzle calculations, 

W = AVd. 

A = 7rDx = leakage area in square feet. 

d = steam density at any row in pounds per cubic foot. 

V = steam velocity through the clearance space in feet per 
second. This velocity is, of course, a function of the pressure 
difference in each row of blades but it should be noted that as 
the available energy utilized in each stage of a group of con- 
stant diameter is approximately the same, the value of V will 



136 THE STEAM TURBINE 

be approximately the same for all rows in a group. With con- 
stant radial clearance and constant V the leakage flow is pro- 
portional to the steam density, which means that the leakage 
of steam at the high-pressure end of a group of blades of constant 
diameter will be greater than at the low-pressure end. There- 
fore the overall efficiency of the first row of blades is lower than 
the rest. It should be noted that the ratio of the radial clear- 
ance x to the blade height gives no correct measure of the ratio 
of leakage steam to useful steam flow. Actually this ratio is 
larger (about 3 or 4 times, depending on the gauging) ; for it is 
obvious that the leakage steam flows axially instead of at an angle 
of about 20 , as is the case for the main flow between blades. 
The steam velocity V through the leakage areas can be deter- 
mined from the available energy utilized per row. The lower 
the available energy per row, the lower the velocity across the 
clearance space and hence the lower the leakage. Hence the 
practice of reducing the diameter of the drum and using a lower 
peripheral velocity in the high-pressure end to suit the lower 
available energy utilized per row is beneficial in reducing leakage. 
This reduction in diameter has the additional effect of reduc- 
ing the leakage area (xD) and hence to further minimize the 
loss due to leakage. 

FORMULA FOR LEAKAGE THROUGH DUMMY RINGS ON BALANCE 

PISTONS.* 

A = leakage area in square inches = dx. 

d = diameter packing ring in inches. 

x = clearance between rings in inches. 
Pi = initial, P 2 final pressure in pounds per square inch 
absolute. 

n = number of rings or labyrinths. 
W = leakage in pounds per second. 
Vi = initial specific volume in cubic feet per pound. 

* See Die Dampfturbinen by Stodola, page 319, 5th Edition. 



STEAM TURBINE TYPES AND BLADE DESIGN 



137 



As a first illustration of a possible design consider a 9000 
horsepower, 2400 revolution per minute, condensing turbine of 
the combined impulse and reaction type. The first expansion 
from an initial steam pressure of 165 pounds per square inch 
absolute to 60 pounds per square inch absolute takes place in 
an impulse wheel having two rows of moving blades of a mean 
diameter of 43 inches or blade speed of 450 feet per second. 
The reaction blading proper consists of an intermediate section 
single flow and a low-pressure section double flow (see Fig. 107). 
The available energy between 60 pounds per square inch abso- 
lute and one pound per square inch is about equally divided 
between the intermediate and the low-pressure sections. The 
blading is designed as follows : 



Intermediate (Single Flow). 


Low-Pressure Drum (Double Flow). 


Mean Diam., 


No. of 


Blade Height, 


Mean Diam., 


No. of 


Blade Height, 


Inches. 


Rows. 


Inches. 


Inches. 


Rows. 


Inches. 


31 


7 


3 


43* 


2 


3, 


3*2 


S 


Ah 


44i 


2 


4* 


35 


4 


7 


47 


2 


7 








49 


2 


9 



* Observe that this value is ^ times the diameter of the first row in the preceding (interme- 
diate) section. See page 113. 

The blade width varies from \ inch for the 3-inch blades to 
1 inch for the 9-inch blades. We have therefore 25 pressure 
stages including the first impulse stage. 

Consider as a second illustration a 1500 kilowatt, 3600 revo- 
lution per minute, non-condensing reaction turbine to operate 
between 160 pounds per square inch absolute initial pressure 
and 30 pounds per square inch final pressure. Steam flow is 
approximately 16 pounds per second. Available energy is 126 
B.T.U. per pound of steam. Consider two sections of reaction 
blading, one high pressure at a blade speed of 250 feet per second 
(16 inches mean diameter) and another section at a blade speed 
of about 375 feet per second (24 inches mean diameter). 

With 20 discharge angles of blades, 70 entrance angles, and 



138 



THE STEAM TURBINE 



.92 velocity coefficient * or .845 energy loss f coefficient we have 
by reference to the formulas on page 92, 



Vi= VbWi7f _„,„ v _^4_ 



F r2 = 



sm 50 
Vb sin 20 



250 X 



= 250 X 



.766 
•341 



307 feet per second, 
in feet per second. 



sm 50" .766 

Available energy E per stage for the high-pressure group with 
a blade speed of 250 feet per second is 
E = (307 2 - in 2 ) -=-(.845X778X32.16) =3.9 B.T.U. per pound. 
The available energy per stage in any other stage will be 
proportional to the square of the blade velocity selected. The 
number of stages or rows will be chosen to utilize the full range 
of available energy as follows: 



No. of 
Rows. 


Mean Diam- 
eter, Inches. 


B.T.U. per 
Stage. 


B.T.U. per Group. 


Specific Volume, 
Cu. Ft. per Lb. 


IO 
8 

2 
2 
2 


16 1 

17 ) 

24 ) 
24-5 ( 

25 ) 


39 

4-4 

8.8 

9-15 
9-5 


3-9 Xio = 39 
4.4 X 8 = 35 
8.8 X 2 = 17.6 
9.15 X 2 = 18.3 
9-5 X 2 = 19 


2.8- 4-3 
4-3" 6.4 
6.4- 8.1 
8.1- 9-7 
9.7-12 .6 




Total 128.9 





In choosing these numbers of rows of blades the method is 
as follows: From the specific volume, as found in the steam 
tables, the annular area for each group is determined. This 
area determines the mean diameter of the group. The B.T.U. 
per stage may then be found, since it is proportional to the 
square of the mean diameter. Finally the number of rows in 
each group is determined so that the B.T.U. for each group is 
about -g 3 - in the high-pressure section and - 6 g 3 - for the low-pressure 
section. Of course experience is the only sure guide in choosing 
dimensions of this kind. 

* A usual assumption for the first approximation. After the steam velocities 
have been determined, a second " corrected" calculation can be made with values 
of coefficients from curve " B " in Fig. 51. 

t Velocity is proportional to the square root of the available energy. In this 
case (.Q2) 2 = .845. Observe that sin 70 = sin no°, in reference to the formulas 
on page 92. 



STEAM TURBINE TYPES AND BLADE DESIGN 



139 



Leakage through Labyrinth Packing. Assuming a radial 
clearance of .02 inch and 30 labyrinths or rings on a 16 inch 
diameter, and using the formula previously given, leakage is 
found to be .65 pound per second. Hence 16 — .65 = 15.35, 
which is the actual flow of steam in pounds per second through 
the turbine proper and available for doing work. 

Steam Leakage Around Tips of Stationary and Moving 
Blades. Assuming a radial clearance of .04 inch and the en- 
ergy division as above we obtain for the first five rows a leakage 

flow of — '- *—*2 = 1-45; say 1.5 pounds per second; where 

144 X 3.06 

2.1 

^— is the leakage area in square feet, 307 is the steam velocity 

*44 

in the leakage space, and 3.06 is the average specific volume in 

cubic feet per pound in the first stages. 

Computing the other leakages in a similar manner we obtain 

the following tabulation: 



No. of Stages, 
per Group. 



Radial 

Leakage, Lb. 

per Sec. 



1-5 

I .2 

I 
O.9 

i-3 

1 



Effective 
Steam Flow, 
Lb. per Sec. 



85* 


19. 


2 


19- 


35 


i7- 


45 


i7- 


05 


17. 


25 


18. 


35 


19 



Output in B.T.U. per Lb. 



5X 
5X 
.6X 
.6X 
6X 
3X 
X 



.845 = 16 

.845 = 16 
.845 = 14 
.845 = 14 
.845 = 14 
•845 = 15 
.845 = 16 



B.T.U. Output 
per Sec. 



228 

234 
212 
214 
208 
222 
23O 



Total 1548 



Total output in kilowatts 



1548 X 3600 _ 
3412 



i635-t 



Or allowing 35 kilowatts for bearing friction loss and other un- 
accounted for losses, we obtain a net shaft output of 150 kilo- 
watts and a steam economy of 36 pounds per shaft kilowatt 
hour thus, 

16 X 3600 _ 
1600 



30 pounds 



* 15.35 - 1.5 = 13.85 pounds. 

t Observe that this result is in kilowatts, and not horsepower as in the preced- 
ing illustrative example. 



I 4 o THE STEAM TURBINE 

This is equivalent to an efficiency of 73 per cent. This ex- 
ample shows a very favorable application for the reaction tur- 
bine in so far that it indicates how a sufficient number of rows 
can be adopted to utilize the available energy efficiently, and 
yet maintain low leakage losses by using a large number of 
stages of comparatively small diameter. 

Reaction Blading for a Turbine of 7500 Kilowatts* Capacity. 
Steam conditions are 180 pounds per square inch absolute, 
i5o°F. superheat and 28 inches vacuum. The type is double 
flow for intermediate and low-pressure blading, and has an im- 
pulse wheel (34 inches diameter) with two rows of blades for the 
high-pressure end. First expansion in impulse wheel to 67 
pounds per square inch absolute, 90 B.T.U. available. Second 
expansion in intermediate drum to 10.5 pounds per square inch 
absolute, 137 B.T.U. available. Steam flow is at rate of 25 
pounds per second from nozzles. Third expansion in low-pres- 
sure drum to 1 pound per square inch absolute, 141 B.T.U. 
available. Allowance was made for reheating between the 
expansions. 

For reaction blading allow a velocity coefficient of .92 or 
energy coefficient .845 in the intermediate section and a velocity 
coefficient of .90 or energy coefficient .81 in the low-pressure 
section, since steam velocities are higher. 

For 20 exit angle the following relations hold as shown in 
previous chapters: Ratio of actual steam exit velocity to periph- 
eral velocity -■ 1.23; ratio of entrance relative velocity to 
peripheral velocity = .445; available energy to be utilized per 
pressure stage. 

E = Hi W . 
/ X 25,000 

/ = friction energy coefficient which is .845 for interme- 
diate and .81 for low-pressure section. 
Vt = blade speed in feet pei second. 






STEAM TURBINE TYPES AND BLADE DESIGN 141 



INTERMEDIATE BLADING (DOUBLE FLOW). 



Mean Diam. 
Inches. 



No. of 
Rows. 


v b , 

Ft. per Sec. 


7 
6 

5 


33° 
346 
375 



Available Energy, 
B.T.U. 



6.8 X 7 =48 
7.5 X 6 = 45 
8-75 X 5 = 44 



Specific Volume, 
Cu. Ft. per Lb. 



7 -11 -4 
11 .4-21 .8 
21.8-34 



Blade 
Height, 
Inches. 



52 



This table is constructed in the same manner as in the pre- 
vious example. Total available energy for intermediate blad- 
ing, 137 B.T.U. per pound. Useful energy per pound of steam 
= 137 X .845 = H5-5 B.T.U. 

LOW-PRESSURE BLADING (DOUBLE FLOW). 



Mean Diam., 
Inches. 


No. of 
Rows. 


v b , 

Ft. per Sec. 


Available Energy, 
B.T.U. 


Specific Vol., 
Cu. Ft. per Lb. 


Blade 
Height, 
Inches. 


34-5 
37 
39 
40 


2 
2 

I 
I 


540 
58o 
6lO 
628 


2 X 19 =38 

2X22 =44 

1 X 24.5 = 28 

I X3I.5 = 315 


34-8- 57 

57-110 

1 10-170 

170-300 


4l 
7 
9 
10 



Total available energy is 141 B.T.U. and useful B.T.U. per 
pound of steam is 141 X .81 = 134. In the last two rows 
higher available energies per stage are allowed than according to 
formula on account of appreciably higher velocities at the tips 
of the blades as compared with the blade velocities at the mean 
diameter. 



IMPULSE ELEMENT. 

The impulse wheel (34 inches diameter) has a blade speed of 
530 feet per second, and the available energy is 90 B.T.U. per 
pound of steam. Nozzle velocity coefficient is .96, and bucket 
velocity coefficient is .85. The useful B.T.U. output per pound 
of steam can be obtained from velocity diagram as follows: 



142 



THE STEAM TURBINE 



Vo 2120 ft. per sec. 90 B.T.U. 

V 1 

v 2 

v t 

v t 

v b 

V, 

V-, 

v s 



2040 

1540 


83 ) , 

47-5 i ^ S 


1310 

845 


; ^ \ 20 

14.2 ) 


715 
465 


10.2 j 6 

4-3 ) 


400 

276 ' 


: s I ■•' 



63.2 " 

Useful B.T.U. per pound of steam = 63.2. 

Leakage is calculated below, assuming radial clearance of 
.035 inch for intermediate reaction blading and .05 to .06 inch 
for low-pressure blading. 

First group intermediate ^ ^ - = .8 lb. per sec. 

144 X 9 * F 

Last group intermediate '~^^z — = .4 lb. per sec. 

* ^ i44 X 16.5 4 p 



First group L.P. 
Last group L.P. 



700 X .05 X 44 „ 

1 — = .48 lb. per sec. 

144 X 70 r 

775 X .06 X 50 „ 

J - L2 ^ = .21 lb. per sec. 

144 X 240 

We can assume a leakage of .5 pound per second at each end 
for the intermediate blading, and .3 pound per second at each 
end for the low-pressure blading. For a steam flow through 
the nozzles of 25 pounds per second (90,000 pounds per hour), 
we have the following as the useful output from the blades, 

From impulse wheel, 

25 X 63.2 X 3600 
2545 
From intermediate blades, 

24 X 115.5 X 3600 ' 
2545 

* Average specific volume, cubic feet per pound. 



= 2230 horsepower. 



3920 horsepower. 






STEAM TURBINE TYPES AND BLADE DESIGN 143 

From low-pressure blades, 

24.3 X 114 X 3600 , 

— ^ - = 3920 horsepower. 

Total 10,070 horsepower. 

Allowing seventy horsepower for rotation loss of impulse 
wheel and for bearing losses we have a net output of 10,000 
horsepower. Hence the steam economy is 90,000 -r- 10,000 or 
9 pounds per horsepower per hour which is equivalent to an 
efficiency of 78 per cent. 

The blade heights were determined from the knowledge of 
steam flow, steam velocity and steam density in each group. 
It is customary, in order to reduce the number of standard 
heights of blade, to maintain a constant height of blade in each 
group and obtain the necessary increase in area for steam ex- 
pansion by " gauging," that is by increasing the normal distance 
between blades. It is of course necessary to compute and 
specify the exit area for each row of stationary and moving 
blades. From the energy division given above the steam veloc- 
ity and specific volume at the exit of each row is known, hence 
the exit areas follow from the simple relation, 

W = > 

144?; 

where 

W = steam flow in pounds per second. 
a = exit area in any row in square inches. 
v = specific volume in cubic feet per pound of steam. 
V = actual steam velocity in feet per second in the blades. 

On account of the comparatively high blade velocities in the 
last rows in the low-pressure drum, it will be necessary to hold 
these by the dovetail or " T-head " construction (page 149) 
usually adopted for impulse blading. This will be necessary in 
order to reduce stresses to allowable values. It will also be 
necessary to use nickel steel or some other steel alloy suitable 
for stresses of about 30,000 pounds per square inch at 20 per 
cent, above rated speed. 



144 THE STEAM TURBINE 

We have previously discussed the meaning of the expression 

2nV 2 , 

where V is the peripheral velocity of the wheel or drum in feet 
per second and n is the corresponding number of moving rows. 
In turbines of modern design of capacities of 2000 kilowatts 
and above, the value of this summation varies between 4 X io 6 
to 4.75 X io 6 for condensing turbines of the combined impulse 
and reaction type. The summation should of course also in- 
clude the impulse blading. 

Purely reaction turbines are now built only in the very large 
sizes, 25,000 kilowatts and above, in which case the summation 
specified above will have higher values of 5 X io 6 to possibly 
6 X io 6 . 

The expression given above can be expressed more conven- 
iently in terms of the mean diameter of blades in inches (D) 
and revolutions per minute (N) as follows: 



x*(-)'(-)'- c - 

** \io/ Vioo/ 



In the same connection purchasers of steam turbines should 
guard well their interests by exercising good business judgment 
in purchases. Like all other kinds of machinery, there will be 
" troubles " with new types of steam turbines, and unless the 
manufacturer is known to be financially responsible and well 
established in the business, the purchaser should not buy until 
he has made very careful investigations of the merits of the new 
machines; and he should always insist on having accurate and 
complete acceptance tests, made preferably by disinterested 
engineers of repute. 

Exercise. — Design the blades for a 300-horsepower (maxi- 
mum output) impulse turbine with two pressure stages and two 
velocity stages in each pressure stage (Curtis type). Initial ad- 
mission pressure is 165 pounds per square inch absolute at ioo° F. 
superheat, and the exhaust pressure is 1 pound per square inch 
absolute. Blade speed 500 feet per second. Reheating factor 
is 50 per cent. Use 8 nozzles and arrange for equal energy dis- 



STEAM TURBINE TYPES AND BLADE DESIGN 145 

tribution in the various stages. Nozzle loss is 2 per cent, of 
velocity, and take blade losses from curves on page 101. 

Exercise. — Design of the blades for a reaction turbine with 
50 stages (Parsons type) for the same conditions of power, pres- 
sures and superheat as in the preceding example. English 
method (pages 120 to 131). 

Exercise. — Design the blades of a combined impulse and 
reaction turbine, having a single pressure stage of the impulse 
type with two velocity stages (Curtis type) and reaction blading 
for intermediate and low-pressure stages. Conditions of power, 
initial pressures and superheat are to be the same as in the pre- 
ceding exercises. Assume the expansion in intermediate sec- 
tion is from 45 pounds per square inch absolute to 5 pounds per 
square inch absolute. 

GENERAL COMPARISON OF COMMERCIAL IMPULSE AND 
REACTION TURBINES. 

IMPULSE. 

i. Few stages. 

2. Expansion in nozzles. 

3. Large drop in pressure in a stage. 

4. Initial steam velocities are in general high (1000 to 4000 feet per second). 

5. Blade velocities 400 to 1200 feet per second. 

6. Best efficiency when blade velocity is nearly half the initial velocity of the 
steam. For a single wheel per pressure stage. 

REACTION. 

1. Many stages. 

2. No nozzles. 

3. Small drop in pressure in a stage. 

4. All steam velocities are low (300 to 600 feet per second). 

5. Blade velocities 150 to 400 feet per second. 

6. Best efficiency when the blade velocity is nearly equal to the highest velocity 
of the steam. 

Radial Blade Clearances. In impulse turbines the radial 
clearance (between the blade ring and the inside of the casing) 
is not important. It is one of the first principles of a good 
design of an impulse turbine that the blades shall be made long 
enough to allow the steam to be discharged through them freely 
without " choking " the flow and " spilling " steam over the 
outer edges of the blades. Since the pressure is the same on the 



146 



THE STEAM TURBINE 



two sides of the blades, radial blade clearances in impulse tur- 
bines can be made of generous dimensions. (See Figs. 57 and 

119, in which Curtis designs are shown.) 




Fig. 57. Illustrates Radial and Axial Clearances in an Impulse Turbine. 

In reaction turbines, on the other hand, it is very necessary 
to make radial clearances as small as is mechanically possible, 
because in these turbines the steam expands in the moving as 
well as in the stationary blades and there is a drop in pressure 
between the two sides of every row of blades. On account of 
this pressure drop there is a continuous flow of steam around the 
edges of the blades, which is large or small in amount in pro- 
portion to the size of the radial clearances. The clearance 
between the stationary blades fixed to the casing and the surface 
of the rotor is of course just as important as that between the 
moving blades and the casing. An American manufacturer of 
the Parsons reaction turbines states that the radial clearances 
are from .02 to .10 inch, varying with the diameter of the drum. 
These limits are given for drums between 1 foot and 10 feet in 
diameter. Radial clearances of large sizes of Parsons turbines 
made by Brown-Boveri & Co. are from 2 to 3 millimeters (.08 
to .12 inch). Attainment of minimum safe radial clearances is 
the goal for every designer of reaction turbines. 

Axial Blade Clearances. Axial clearances with respect to 
impulse and reaction turbines present conditions just opposite 



STEAM TURBINE TYPES AND BLADE DESIGN 147 

from those for radial clearances. In reaction turbines, axial 
clearance is not an important factor in the design. Until re- 
cently, however, it was considered very important in the design 
of impulse turbines to make the axial clearance between the 
moving blades, and nozzles or stationary blades, as small as pos- 
sible; and indeed, unfortunately, some impulse turbines in the 
early days were built with very small axial clearances, so that 
the least vibration of the shaft would cause striking of the mov- 
ing blades against the nozzles. It has been shown, however, by 
actual experience as well as by experiment that axial clearances 
can be made as large as .20 inch without appreciable loss; or, in 
other words, practically as large as in reaction turbines — usu- 
ally about .10 to .20 inch. 

The difficulties of the designers of the first commercial im- 
pulse turbines can well be imagined when it was considered so 
essential to make the axial clearances not more than .02 or .03 
inch. In the case of one small turbine built with three stages 
the axial expansion of the shaft in the length included between 
the high-pressure nozzle mouths and the blades of the third 
stage was .10 inch by actual measurement. To allow for a shift- 
ing of the blades of .10 inch with only .03 inch axial clearance in 
a turbine with velocity stages was not an easy problem. 

Axial clearances in Curtis impulse turbines are .06 to .15 inch 
for 500-kilowatt sizes, and in larger machines are sometimes as 
much as .25 inch. In Rateau impulse turbines these clearances 
are from .12 to .25 inch.* 

Materials for Blades and Erosion. Ordinary rolled steel is a 
very suitable metal for turbine blades when used for dry or super- 
heated steam at either high or low velocities if the turbine is 
kept in practically continuous operation. Wet steam, however, 
will wear away steel blades very rapidly by erosion, and when a 
turbine fitted with steel blades is idle for days at a time the 
blades will corrode, so that when it is started again the particles 

* In impulse turbines with nozzles discharging radially into blades or buckets 
on the rim like the Sturtevant, Terry, or Riedler-Stumpf types, it is stated that there 
is no appreciable change in velocity loss when the radial clearance (between the 
nozzle and the buckets) is increased from .10 to .40 inch. 



148 



THE STEAM TURBINE 



of iron oxide (rust) will be carried away by the steam to act 
like a sand blast on the blades in succeeding stages. Steel is an 
exceptionally good material for blades under favorable condi- 
tions because it can be rolled cheaply into bars of any shape of 

section,* and it is unequaled 
for strength. Copper alloys, 
known in the trades as " ex- 
truded metal," are made into 
bars of any shape of section 
by " drawing " as wire is 
manufactured. Blades of this 
material are not strong enough 
for the highest velocities used 
in some modern types. 

No metal has all the phys- 
ical properties desirable in a blading material. A compound 
metal known as Monnot or "duplex" metal has been developed 
but has found little use, which is a disappointment. It consists 
of a steel core covered with a thin copper sheathing chemically 




Fig. 58. Etched Section of a Blade made 
of Monnot Metal (Steel and Copper). 






Fig. 59. Fig. 60. Fig. 61. 

Designs of Steam Turbine Blades. 

welded to the steel in such a perfect manner that the blades may 
be drawn cold from the original ingot into the required finished 
section without in any way affecting the bond between the cop- 
per and the steel. Fig. 58 shows an etched section of a blade 
of this material from a Westinghouse turbine. 

+ Rolled bars are cut up into lengths corresponding to the height of the blade 
plus an additional length for dovetailing into the rim of the turbine wheel. When 
this dovetailing method is used (Fig. 63) the blades are separated from each other 
by " spacing pieces " of suitable shape to fit between the blades. 



STEAM TURBINE TYPES AND BLADE DESIGN 149 



Blades like those, for example, in Figs. 59-61, which are too 
irregular to be rolled or drawn are usually drop-forged, or if very 
irregular and for low velocities they may be cast of bronze or 
copper alloys. Forked blades (Fig. 59) are commonly cast with 
the forks far enough apart so that they will pass over the en- 
larged section of the rim and are forced together when they are 
in place. Another method is to cut away the enlarged part of 
the rim section for a short length, and blades drop-forged or cast 
with the forks in their normal position can be inserted at this 
place and can then be pushed around on the rim till all the blades 
are in place. The parts of the rim cut away must be replaced 
to secure the blades at that section. 

The blades of small sizes of Curtis turbines are sometimes cut 
in the rim of a solid disk by automatic machinery. De Laval 





60' 



Fig. 63. Dovetailed Type of Blade 
(Curtis). 



Fig. 64. Typical De Laval Blades- 



i5° 



THE STEAM TURBINE 



blades are made of an alloy of steel containing nickel and copper. 
The metal is drop-forged into the peculiar shape required for in- 
sertion into the blade disk. (See Fig. 64.) These blades have 
a smooth, hard, glossy coating of natural oxide which is said to 
have excellent properties of resisting the action of water, super- 
heated steam, and corrosive acids. On the other hand, when 
" mild " nickel steel blades are used there is much trouble from 
rust, and " high " nickel steel blades are too brittle. 

It is stated that the usual alloy used in England for blades of 
Parsons turbines is 63 Cu + 37 Zn; but any zinc alloy is quite 
unsuitable for superheated steam or for high velocities. 

Fig. 65 * shows the effect of the erosion due to steam on blades 




Fig. 65. Photograph of Turbine Blades Showing Erosion. 



made of Delta metal about 60 Cu + 37 Zn + 3 Fe. These 
blades were held stationary in a steam jet for 128 hours. The 
blades on the left side of the figure were subjected to steam at 
2900 feet per second; and those on the right to steam at 600 
feet per second. Low-velocity steam eroded the blades so little 
that the tool marks put in the blades when they were made are 
still visible. 

* The author is indebted to Mr. Francis Hodgkinson for this photograph. 



CHAPTER V. 

MECHANICAL LOSSES IN TURBINES. 

In the designs of turbines on the preceding pages the nozzle 
and blade efficiency was first calculated, and then the total, or 
"over-all," shaft efficiency was obtained by subtracting other 
losses as follows: 

(i) Disk and blade friction, or windage, due to rotation in a 
fluid medium (steam). 

(2) Leakage of the steam chiefly through the clearance 
between the shaft and the diaphragms (" stage leakage ") of a 
multi-stage impulse turbine and through the radial blade clear- 
ances in a reaction turbine. 

(3) Bearing and stuffing-box friction losses 

(4) Radiation. 

Of these, the first three are, in a way, mechanical losses in the 
sense that the details of mechanical design largely determine 
their values. 

The first of these losses, disk and blade rotation loss, is by far 
the most important and will be discussed first. 

Losses Due to Friction of Turbine Wheel Revolving in Steam. 
Losses due to revolving disks or wheels in steam are very diffi- 
cult to determine with accuracy. Tests to determine these 
losses are usually made with the wheel rotating in stagnant 
steam, and it is practically impossible to have, under these con- 
ditions, steam of the same quality or superheat in all parts of the 
casing. A number of formulas have been proposed for the 
friction losses of disks and blades in dry saturated steam, but 
there is no good agreement of the results of different experi- 
menters. In fact no great accuracy can be expected because 
there is no doubt that the exponents of logarithmic friction 

151 



152 THE STEAM TURBINE 

curves plotted from such tests vary considerably with the details 
of design, and besides, it is very difficult to get good tests.* 

An important reason why the tests from different designs of 
turbines do not agree better is that clearances between moving 
and stationary parts have an appreciable effect. If the clear- 
ances all around the wheel are very small the wheel and blade 
friction loss will be somewhat less than for a wheel revolving in 
large clearance spaces. This effect is most marked at low speeds. 
When higher speeds are reached there is more tendency for the 
wheel to "cut through" the surrounding steam without increas- 
ing the "disturbance" in proportion to the increase in speed. 

The author has from time to time investigated large numbers 

of tests to determine the friction losses of wheels and blades of 

turbines in steam and air, and this experience has shown that the 

following formulas will give fair average results for forward 

running in practically stagnant steam. The rotation loss or 

skin friction of a plain disk f revolving in dry saturated steam is 

expressed by the following formula in horsepower : 

/ 11 \ 2 - 
F,„ = .08 d 



G™) 7 > (3 ° a) 



where d is the diameter of disk to inner edge of blade in feet, 
u is the peripheral velocity of disk in feet per second.}: 
y is the density of surrounding medium in pounds per 
cubic foot (reciprocal of the specific volume). 
A similar term to determine the rotation loss of one row of 
blades F & (without the disk), in horsepower, is 

F» =. 3 dl'- 5 (— Y'V (30b) 

\I00/ 

* The peculiar circumstance that water in the liquid state can exist, almost 
indefinitely, in the presence of superheated steam, leading some to propose a 
vergasungswarme, is one of the greatest difficulties. 

t Similar to those in Curtis and Rateau turbines. On account of the thick 
hubs of De Laval disks (Figs. 83 and 84), about 15 percent, should be added to 
the results given by equation (30a) to allow for the larger surface of these disks. 

| It is often stated that the disk and blade friction losses vary as the third power 
of the speed. But this value cannot be stated with any claim to great accuracy. 
Experimenters do not all agree on this value, and values from 2.5 to 3.5 are given 
by different authorities. The author, from the result of the experiments he has 



MECHANICAL LOSSES IN TURBINES 



153 



where 1 = length of blades in inches excluding the band (if 
there is one), and d, u, and 7 are used as before. 

For a simple turbine wheel with only one row of blades we 
can write for the total rotation loss ¥ t in horsepower : 

F t = (.08 d + .3 F) \ — 1 ' 8 d T . (30c) 

( 100 ) 

The density of superheated steam varies with the amount of 
superheat, so that by adding the following notation, 
y d = density of dry saturated steam at the pressure of the 

surrounding medium in pounds per cubic foot, 
D = superheat in degrees F., 

v d = specific volume of dry saturated steam at the pressure in 
the surrounding medium in pounds per cubic foot, 
and using the following equation for specific volume v 8 of super- 
heated steam given on page 40, 

v. = (1 + .00065 D) 2 v d , 
we have the following formulas, taking the place of (30a), (30b), 
and (30c) above, for superheated steam: 

r»=. 3 dl" \ 100 ' . (aoe) 



(1 + -00065 D) 2 

\I00/ 



f, = (.08 d + .3 n^r^-—-. 30f) 

(1 + .00065 D) 2 

investigated, considers the 2.8 power a good average value suitable for practically all 
conditions. In the value of the exponent this rotation loss resembles train and 
ship resistance. The windage loss of dynamos properly designed for high speeds is 
a curve of the second power. When the windage loss curve of a dynamo shows an 
exponent of 3 or 3.5 it must be inferred that the machine was not properly designed 
for high speeds. It may be interesting to the practical men reading this book to know 
how the exponent is obtained from a test. This is done most conveniently by 
plotting on any suitable coordinate paper the logarithms of the loss for the ordinates, 
and the logarithms of the speed for the abscissas. The tangent of the curve is the 
value of the exponent if the scales of ordinates and of abscissas are the same. 



i54 



THE STEAM TURBINE 



Or the curve given in Fig. 68 can be used to correct equations 
(30a), (30b), and (30c) by means of a coefficient. 

While the effect of superheating is to reduce these losses, 
moisture, on the other hand, increases them very appreciably. 



' 


























































~1.00 


















































































































.2 .90 
o 

& 
6 -80 


































































































































































































































70 



























































2.60 



2.40 



2.20 



2.00 



§ 1.60 

a 

1.40 



1.20 



1.00 



40 60 



100 120 140 160 180 200 220 
Superheat-Degrees Fahr. 



Fig. 68. Curve to Correct Rotation Losses for Superheat 



r 


T^ 


_~f 


/ 


7 


7 


2 


7 


1 ~7 


~7 


7* 


^ 


•^ 


^' 


^^ 


^^ 





10 12 14 16 18 
Percent. Moisture in Steam 



22 24 26 



Fig. 6q. Curve to Correct Rotation Losses for Moisture. 



Fig. 69 shows a curve giving the coefficients to be applied to the 
losses calculated by the above formulas for dry saturated steam 
to correct for moisture. 

Example, (a) Calculate the frictional rotation loss of a disk 
3 feet in diameter of a non-condensing single stage turbine (steam 



MECHANICAL LOSSES IN TURBINES 155 

pressure 15 pounds per square inch absolute) when the steam 
is (1) dry saturated, (2) superheated ioo° F., (3) 10 per cent 
wet. The speed is 3600 revolutions per minute, (b) Determine 
also the rotation loss of a single row of blades 1 inch long on this 
disk. Ans. (a) 3.50; 3.06; 4.38. (b) 4.37; 3.82; 5.47. 

At high peripheral speeds the rotation loss of a non-condensing 
turbine with the wheels revolving in steam at atmospheric pres- 
sure is quite large, as the example above illustrates. This loss 
decreases, however, very rapidly with increasing vacuum, and 
is, in fact, nearly proportional to the pressure. This fact is not, 
however, always appreciated by designers. Of course, when 
disk and blade rotation losses are being calculated for a series 
of pressures for the several stages of a turbine, as is usually 
done before deciding on the nozzle proportions, it is only neces- 
sary, if the wheel dimensions are constant, to calculate for one 
pressure and determine the values for the other stages by multi- 
plying by a constant representing the ratio of the densities. Of 
all the variables in equations (30a), (30b), and (30c), the density 
is the only term varying as the first power. For most work it 
will be allowable to assume, within a small range, the density pro- 
portional to the pressure ; that is, if the disk and blade loss has 
been calculated in steam at some given pressure, the correspond- 
ing friction loss at any other pressure may be found by the ratio 
of the pressures. 

The disk and blade rotation losses of a Parsons or other drum 
type may be calculated with the above formulas by calcu- 
lating the loss for each group of blades of the same length and 
diameter and adding to the sum of the blade losses the rotation 
loss due to disks approximately equivalent to the outside surface 
of the drum. As the friction loss due to the drum itself is small 
compared with that of the many rows of blades, no great accuracy 
need be attempted in this calculation. 

In small sizes of steam turbine-generators the rotation loss is a 
considerable percentage of the total output. The disk and blade 
loss of a single stage turbine with a single row of blades, rated by 
the manufacturer at about 250 kilowatts at 3600 r.p.m., is shown 



i56 



THE STEAM TURBINE 



in Fig. 70. The curves show that the rotation or windage loss 
of the generator alone is about 30 kilowatts and the total rotation 
loss is 50 kilowatts or 20 per cent, of the rated output. Similarly 



70 



60 



50 



40 



10 











































/' 
























































































































































































































& 










































C4 


/ 








































-^ v 


sp> 








































c 


w 


J^ 








































if. 


v^ 






































rfO 




&*% 



















iS-- 


































SOW 


uoti 


v> 


&z 
































)\SC- 


&oA 


B^ a 


dej 








Curves are Corrected 
For I 2 It Loss 




"""' 










V 































































2000 



2500 



3000 
Speed-R.P.M. 



3500 



4000 



Fig. 70. Rotation Loss Curves of 250-Kilowatt Turbine-Generator. 



the total rotation loss of a 2000 to 3000 kilowatt turbine-gen- 
erator is from 10 to 15 per cent, of the rated output. 

Method of Making Tests to Determine Wheel and Blade 
Rotation Losses of a Steam Turbine. The simplest method for 
making such a test, and the one commonly employed, is to attach 
an electric motor to the turbine shaft (sometimes in a direct- 
connected set the generator is used as a motor) and run it at a 
number of different speeds. In taking a series of speeds, no 
observations are made until conditions have become "steady," 
and the speed must be held constant for several minutes so that 
a number of readings can be taken on the electrical instruments 
measuring the input of the motor. The results give the rota- 
tion loss of the wheel and blades in steam as well as bearing 
friction and the rotation or "windage" and electrical losses of the 
motor. Then the turbine wheel is removed, leaving the packing 



MECHANICAL LOSSES IN TURBINES 157 

at the generator end of the turbine on the shaft, and the motor is 
run alone. The power now measured is that required to over- 
come the rotation and electrical losses of the generator and the 
bearing friction. Curves of power and speed as variables (Fig. 
70) are plotted for each set of observations, and the disk and 
blade loss is determined by subtracting the ordinates of one curve 
from those of the other. It may be assumed with sufficient 
certainty that the weight of the turbine wheel itself would not 
alter the bearing losses to any considerable extent.* 

The important fact that all results given here are for disks and 
blades revolving in a stagnant medium must not be overlooked, 
and it must not be assumed that the results will be the same 
under actual operating conditions. It may be a coincidence that 
the losses are the same in both cases. Under operating condi- 
tions, the spaces between the wheel blades are filled with steam 
flowing from the nozzle over the blades and then to the condenser. 
Now it has been shown by a series of experiments by Laschef 
of the Allgemeine Electricitat Gesellschaft (Berlin) that increasing 
the number of nozzles around the turbine wheel reduces the disk 
and blade rotation losses. These losses in the blades are very 
largely due to the fan action of the blades which start currents of 
steam just as a centrifugal fan does. In other words, this is what 
Stodola calls "ventilation." With steam flowing through the 
blades, this fan action is largely prevented and the losses are con- 
sequently reduced. Another reason why the disk and blade 
rotation losses should be less when the turbine is operating than 
they are in stagnant steam, is that they are really friction losses, 
or a conversion of kinetic energy into heat, with the effect of 
either superheating or drying the steam. In a turbine with 
more than one stage a part of the heat energy gained as the result 
of the friction is converted in the next expansion into kinetic 
energy or velocity. It is usually assumed that about 15 per cent. 

* It may be interesting to observe that since disk and blade friction is pro- 
portional to the density of the medium, the friction is therefore greater in air than 
in dry saturated steam at atmospheric pressure. This is shown by experiments 
published by Lewicki in Zeit. Verein deutscher Ingenieure, March 28, 1903. 

f Stodola, Die Dampfturbinen, third edition, page 130. 



158 THE STEAM TURBINE 

of the disk and blade losses are regained by the reheating, and 
that therefore the actual friction losses in an operating turbine 
are about this amount smaller than in stagnant steam. In cases 
of full admission true blade friction disappears; and a proportion- 
ate reduction will also take place, according to the degree of 
admission, when it is partial. 

Investigation of wheel and blade friction losses by the author, 
using a modification of the method first suggested by Lasche of 
Berlin, did not show the reduction in these losses to be expected 
when determined under operating conditions. These results, 
however, cannot be considered conclusive, as the type of machine 
used was not well suited for the purpose, and only 25 per cent, of 
the blades were filled with steam. It has been stated that when 
a large quantity of steam passes into the casing through a suit- 
able opening without passing through nozzles and escapes through 
the exhaust (without increasing the pressure), the disk and blade 
rotation losses are increased as much as 20 per cent. This 
apparently is an influence to counteract the effect of filling the 
blades. 

In all the analysis that has preceded there are so many uncer- 
tain variables entering that it is impossible to get agreement, 
although, apparently, we have a large amount of data from which 
to draw. It may be stated, however, that all in all, the best data 
on disk and blade friction seem to show that it is smaller and of 
less significance than the results of most investigators would 
show. 

A little space should be given to Lasche's very interesting 
method.* A turbine-generator set was used in which the number 
of nozzles discharging into the turbine could be regulated and the 
output of the generator was observed for each setting of valves, 
and tests with varying loads were made at a number of different 
speeds. The turbine wheel was then removed from the shaft, 
and by running the generator as a motor the friction losses in the 
stuffing-box at the generator end of the turbine and in the bearings, 
as well as the windage loss of the generator, were determined. 

* Stodola, Die Dampfturbinen, third edition, page 131. 



MECHANICAL LOSSES IN TURBINES 



159 



The resistance of the armature and brushes was also measured to 
calculate the heating (I 2 r) loss. The sum of these losses was cal- 
culated for a number of loads (kilowatts) and curves similar to 
those in Fig. 70 were obtained. Curve A in Fig. 71 shows the elec- 















































































































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a,r 


on 


tor- 






















1 


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O 160 












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ut 










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60 


















3500 1 


.p. 


m. 










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_r 










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± 












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r 










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3* 

£3-20 


/ 




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oz 


3 

zle 


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3p 


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3500 r^Jm- 


340 














































?in? 




























// /''•' 


























$// 


ft 


























































ll 










280 


















1 


I 




























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2C0 


















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I 240 
















/ 


1 


























// 






























/ 






























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& 200 
| 


























































5 180 








































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O 160 












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■w 












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02 ^ 


























































'120 


























































100 








// 






























// 






















80 








// 




























/ 


/ 






















60 






// 






























// 
























40 






/ 


























/// 


























20 


1 







































































' 














•0 M 
2 i 
« J -20 

1 s 

IS 
P 3 


1 


00 




1. 


I 


1 

T 02 


Zl 


i 
IS, 


Dp 


a 




t 







Fig. 71. 



Fig. 72. 



Curves for Determining Disk and Blade Rotation Losses at Operating Conditions. 

trical output at 3500 r.p.m. Curve B in the same figure, represent- 
ing the power delivered to the shaft by the turbine, was obtained 
by adding to the generator output for each set of nozzles open 
(curve A) the corresponding generator losses (windage, heating, 
and bearing friction). The lower portions of curves A and B 



160 THE STEAM TURBINE 

are practically straight lines, and by producing curve B to the 
horizontal axis, its intersection represents on the scale of abscissas 
the disk and blade rotation losses of the turbine at the speed of 
the test and under actual operating conditions. 

By making a series of such tests at different speeds curves of 
rotation losses can be made. Fig. 72 shows typical curves of shaft 
output for speeds of 3000, 3500, and 4000 r.p.m. Although 
this method requires very careful experimenting, the same must 
be said of any other method of obtaining these losses, and most 
of the results that have been published are very poor. At least 
it must be admitted that by this method a number of uncertain 
factors to be considered in the "stagnant steam" method are 
eliminated. 

The lines in Fig. 72 are really the same as "Willans lines ,, and 
might just as well be plotted for total "flow" of steam per hour 
as for nozzles open. In fact in turbines where there are no 
nozzles the "flow" of steam must be used. It is obvious that 
any load curve of brake horsepower giving the total steam con- 
sumption can be used to determine the rotation loss by producing 
the "flow" line to the axis on which the output is scaled. A 
good check on the results of such rotation loss tests is secured by 
observing whether the lines for the speeds near the rated speed 
cross each other at about the rated output. In a good design the 
speed-output curve will be like the curve in Fig. 80, giving 
nearly the same output at speeds considerably above or below 
the rating. 

The no load steam consumptions of 2000, 5000 and 9000 kilo- 
watt Curtis turbine-generators are respectively about 14, 12.5, 
and 8 per cent, of that at full load. In other words these 
percentages are only from one to two per cent, greater than the 
sum of the disk and blade rotation and generator windage 
losses. Generator windage loss is probably about equal to 
the sum of all the turbine losses. It is generally assumed that 
the no load steam consumption of a Parsons turbine (without 
the generator) is about 12 per cent, of that at the normal 
maximum output. 



MECHANICAL LOSSES IN TURBINES 161 

It is stated* that at no load the steam required for very large 
reciprocating engines and generators is probably in no case less 
than 15 per cent, of that used at full load. 

Leakage Loss. The other important mechanical loss in a 
steam turbine is that due to the leakage of steam through the 
passages of the turbine without doing work. In impulse turbines 
of more than one stage this loss is chiefly caused by the leakage of 
steam between the shaft and the diaphragms, In a great many 
turbines no satisfactory packing is provided at these places and 
the loss is sometimes more than 10 per cent, of the total amount 
of steam supplied to the turbine. In reaction turbines the loss 
is due to leakage through the radial clearance passages and is 
large or small in proportion to the size of these clearances. The 
loss is usually assumed to be about 5 per cent, in good Parsons 
turbines. 

Future improvements in the economy of all types of steam tur- 
bines will depend largely on the success of designers in reducing 
these leakage losses. For impulse turbines an improved design 
has been patented by Wilkinson (page 255). In reaction tur- 
bines it can be reduced by making a shorter and stiffer shaft. 

Bearing Friction. This loss is due to the friction of the shaft 
in its bearings, and in a De Laval turbine the friction of the gears 
is usually included. An analysis of the losses in a De Laval 
turbine is given on page 194, where the bearing friction loss is 
given as one per cent. Bearing friction is also discussed in the 
footnote on page 108. 

* Kruesi, Proc. Am. Street and Interurban Railway Engineering Association, 
1907. 



CHAPTER VI. 
METHOD FOR CORRECTING STEAM TURBINE TESTS. 

Standard Conditions for Steam Turbine Tests. If tests of steam 
turbines could always be made at some standard vacuum, super- 
heat, and admission pressure, then turbines of the same size and 
of the same type could be readily compared, and an engineer 
could determine without any calculations which of two turbines 
was more economical for at least these standard conditions. But 
steam turbines and engines even of the same make are not often 
designed for and operated at any standard conditions, so that a 
direct comparison of steam consumptions has usually no signifi- 
cance. 

It will be shown now how good comparisons of different tests 
can be made by a little calculation involving the reducing of the 
results obtained for varying conditions to assumed standard 
conditions. The method given here is that generally used by 
manufacturers for comparing different tests on the same turbine 
(a " checking " process) or on different types to determine the 
relative performance. To illustrate the method by an applica- 
tion, a comparatively simple test will first be discussed. 

Practical Example. Corrections for Full Load Tests. The 
curve in Fig. 73 shows the steam consumption for varying 
loads obtained from tests of a 125-kilowatt steam turbine 
operating at 27.5 inches vacuum, 50 F. superheat, and 175 
pounds per square inch absolute admission pressure (at the 
nozzles). It is desired to find the equivalent steam consump- 
tion at 28 inches vacuum, o° F. superheat, and 165 pounds per 
square inch absolute admission pressure for comparison with 
"guarantee tests" (Fig. 74) of a steam engine of about the 
same capacity operating at the latter conditions of vacuum, 
superheat, and pressure. The manufacturers of the steam 

162 



METHOD FOR CORRECTING STEAM TURBINE TESTS 163 









































































27.5 Ins. Vacuum 

50°F Superheat 

175 Lbs.Per Sq.In.Abs. 

Pressure 





















































































































































































































































































































































































































































































































20 40 60 80 100 120 140 160 180 200 

Output in Kilowatts 

Fig. 73. Load Curve of a Typical 125-Kilowatt Steam Turbine. 































1 
















\ 


























28 Ins. Vacuum 

0°F Superheat 

165 Lbs.Per Sq.In.Abs. 

Pressure 








\ 

\ 




























> 


\A 








































































SB 




V 










































s 














































T^ 1 - 








^ 








_-• 




















































































































A. Steam Consumption of Engine 
























OfS 


-tea 


aiT 


urbi 


ne. 





















20 



40 



GO 



80 100 120 

Output in Kilowatts 



140 



160 



180 



200 



Fig. 74. Comparative Load Curves of a Reciprocating Steam Engine and a Steam 
Turbine — Both of 125 Kilowatts Capacity at Full Load 



1 64 THE STEAM TURBINE 

turbine have provided the curves in Figs. 75, 76, and 77 
showing the change of economy with varying vacuum, superheat, 
and pressure. With the help of these correction curves, the 
steam consumption of the turbine can be reduced to the conditions 
of the engine tests. Fig. 75 shows that between 27 and 28 inches 
vacuum a difference of one inch changes the steam consumption 
1.0 pound. Fig. 76 shows a change of 2.0 pounds per ioo° F. 
superheat, and from Fig. 77 we observe a change of 5.0 pounds 
in the steam consumption for 100 pounds difference in admission 
pressure. Compared with the engine tests the steam turbine was 
operated at .5 inch lower vacuum, 50 F. higher superheat, and 
10 pounds higher pressure. At the conditions of the engine 
tests, then, the steam consumption of the steam turbine should be 
reduced .5 pound to give the equivalent at 28 inches vacuum, 
but is increased 1.0 pound to correspond to o° F. superheat, and 
.5 pound more to bring it to 165 pounds absolute admission 
pressure. The full load steam consumption for the steam tur- 
bine at the conditions required for the comparison is, therefore, 
24.5 - .5 + 1.0 + .5, or 25.5 pounds.* 

Persons who are not very familiar with the method of making 
these corrections will be liable to make mistakes by not knowing 
whether a correction is to be added or subtracted. A little 
thinking before writing down the result should, however, prevent 
such errors. When the performance at a given vacuum is to be 
corrected to a condition of higher vacuum the correction must be 
subtracted because obviously the steam consumption is reduced 
by operating at a higher vacuum. When the steam consumption 
with superheated steam is to be determined in its equivalent cf 
dry saturated steam (o° superheat) the correction must be added 
because with lower superheat there is less heat energy in the 
steam and consequently there is a larger consumption. Usual 

* The corrected steam consumption is found to be nearly the same as that which 
the three correction curves show for the same conditions, that is, about 25.0 pounds. 
If there had been a difference of more than about 5 per cent, between the corrected 
steam consumption and that of the correction curves for the same conditions, the 
" ratio " method as explained on page 130 for fractional loads should have been used 
also for full load. 



METHOD FOR CORRECTING STEAM TURBINE TESTS 16 



£5 30 

aw 

SM 25 
£* 

1 5 20 



15 











































































0°F.Superheat 

165 Lbs.Per Sq.In.Abs. 

Full Load( 125 Kw.) 





















































































































































































































































































































































29 



20 21. 22 23 .24 25 26 27 28 

Vacuum JTnches oOlercury 

Fig. 75. Vacuum Correction Curve for a 125-Kilowatt Steam Turbine 















































30 






























28 Ins.Vacuum 

165 Lbs. Per Sq.In.Abs. 

Full Load(125 Kw.) 


































•2s25 

P 


























































































a & 
c*20 

O u 


























































































|Sl5 
CO 


























































































10 















































20 



60 



100 120 140 160 130 200 

Superbeat-Degs.Fahr. 
FlG. 76. Superheat Correction Curve of a 125-Kilowatt Steam Turbine. 



2 s 

C n 

o » 
fl5 

















































30 


























































































25 


























































































20 


























































































IS 






























28 Ins.Vacuum 

0°F. Superheat 

FullLoad(125Kw.) 
































m 















































100 110 



120 



190 



200 



130 140 150 160 170 180 
Steam Pressure-Lbs. Per Sq.In.Abs, 
Fig. 77, Pressure Correction Curve of a 125-Kilowatt Steam Turbine 



1 66 THE STEAM TURBINE 

corrections for differences in admission pressure are not large; 
but it is well established that the economy is improved by 
increasing the pressure. 

Corrections for Fractional Loads. It is the general experience 
of steam turbine manufacturers that full load correction curves, 
if used by the following "ratio" or percentage method, can be 
used for correcting fractional or over loads. This statement 
applies at least without appreciable error from half to one and a 
half load, and is the only practicable method for quarter load as 
well.* Stated in a few words, it is assumed then that the steam 
consumption at fractional loads is changed by the same percent- 
age, as at full load, for an inch of vacuum, a degree of super- 
heat, or a pound of pressure. It will now be shown how this 
method applies to the correction of the steam consumption of 
the turbine at fractional loads. Now according to the curve in 
Fig. 75 the steam consumption at 27.5 inches (25.6 pounds) must 

2 c o 
obviously be multiplied by the ratio -^— ,t of which the numer- 

25.6 

ator is the steam consumption at 28 inches and the denominator 

at 27.5 inches, to get the equivalent consumption at 28 inches 

vacuum. This reasoning establishes the proper method for 

making corrections; that is, that the base for the percentage 

(denominator of the fraction) must be the steam consumption at 

the condition to which the correction is to be applied. J Similarly 

the correction ratio to change the consumption at 50 F. super- 

2^0 
heat to o° F. is^i—, and to correct 175 pounds pressure to 165 
24.0 

* A very exhaustive investigation of this has been made by T. Stevens and H. M. 
Hobart which is reported in Engineering, March 2, 1906. 

f Assuming that this short length of the curve may be taken for a straight line 
without appreciable error. 

% In nearly all books touching this subject so important to the practical, 
consulting, or sales engineer, the alternative method of taking the steam consump- 
tion at the required conditions as the base for the percentage calculations is 
implied. By such a method percentage correction curves derived from straight 
lines like Figs. 76 and 77 would be straight lines and, in application, give ab- 
surd results. Actually such percentage corrections will fall on curves (see Figs. 87 
and 88). 



METHOD FOR CORRECTING STEAM TURBINE TESTS 167 

24.8 



pounds the ratio is 



Data and calculated results obtained 



24.3 



by this method may then be tabulated as follows: 





Conditions 
of Test. 


Required 
Conditions . 


Correction 
Ratio. 


Percentage 
Correction. 




2 7-5 
5°- 

i75- 


28 



165 


25.O 
25.6 

2 5-° 
24.0 

24.8 

2 4-3 


-2-34%* 
+ 4-i7% 

+ 2.06% 
+ 3-89% 


Superheat, degrees F 


Admission pressure, pounds absolute . . 











* Steps in the calculation are omitted in the table, thus ■ = .9766 or 97.66 per cent., making the 

2S.6 
correction 100 — 97.66, or 2.34 per cent. It may seem unreasonable to the reader that these percent- 
ages are calculated to three figures when the third figure of the values of steam consumption is 
doubtful . In practice, however, the ruling of the curve sheets must be much finer and to larger scale 
so that the curves can be read more accurately. 

The signs + and — are used in the percentage column to 
indicate whether the correction will increase or decrease the 
steam consumption. " Net correction " is the algebraic sum of 
the quantities in the last column. 

The following table gives the results of applying the above 
"net correction" to fractional loads. 



Steam consumption from 

test (Fig. 73), lbs 

Net correction + 3 . 89 %. . , 
Corrected steam consump- 
tion 



■j- Load 
(31.3 kw.) 



31.2 

+ 1.2 

32-4 



\ Load 
(62.5 kw.) 


f Load 
(93.8 kw.) 


^ Load 
(125 kw.) 


26.9 
+ 1.1 


25.2 
+ 1.0 


24-5 

+ 1.0 


28.0 


26. 2 


2 5-5 



(156.3 kw.) 



23.6 
+ -9 



24-5 



Curve B in Fig. 74 shows the corrected curve of steam con- 
sumption for the steam turbine as plotted from the above table. 
By thus combining, on the same curve sheet, curves A and B as 
in this figure, the points of better economy of the turbine are 
readily understood. 



i68 



THE STEAM TURBINE 



Results of economy tests of the various turbines given on the 
preceding pages are of very little value for comparison when the 
steam consumptions or " water rates " are given for all sorts of con- 
ditions. With the assistance, however, of curves like those shown 
in Figs. 75, 76, and 77, if they are representative of the type and 
size of turbine tested, it is possible to make valuable compari- 
sons between two or more different turbines. Some very recent 
data of Curtis and Westinghouse-Parsons turbines are given 
below, together with suitable corrections adopted by the manu- 
facturers for similar machines. 





















































































































X 
























18 


























































































gW.16 


































































' 
























AM 
j^ 

°8.14 






































N 








12 








































\ 












































1 


\ 














































, 


10 















































20 



21 



22 



27 



29 



3 24 25 2 

Vacuum Inches of Mercury 
Fig. 78. Typical Vacuum Correction Curve of a 5000-Kilowatt Impulse 

Turbine. 

The following test of a Westinghouse-Parsons turbine, rated 
at 7500 kilowatts, was taken at Waterside Station No. 2 of the 
New York Edison Co., and a comparison is made with a test 
of a five-stage 9000-kilowatt Curtis turbine at the Fisk Street 
Station of the Commonwealth Electric Company of Chicago. 
As no pressure correction is given for the Curtis machine, the 
New York Edison test is corrected to the pressure at which the 
other machine was operated (179 pounds per square inch gauge). 
Approximately an average vacuum for the two tests is taken 
for the standard, and ioo° F. superheat is used for comparing 
the superheats. These assumed standard conditions make the 



METHOD FOR CORRECTING STEAM TURBINE TESTS 169 

corrections for each turbine comparatively small. When two 
tests are to be compared, by far the more intelligent results are 
obtained if each is corrected to the average conditions of the 
two tests, rather than correcting one test to the conditions of the 
other. There is always a chance for various errors when large 
corrections must be made. 

7500-KILOWATT WESTINGHOUSE-PARSONS TURBINE, WATERSIDE 
STATION NO. 2, NEW YORK EDISON COMPANY. 



Duration of test, hours 

Speed, revolutions per minute 

Average steam pressure, pounds gauge 

Average vacuum, inches (referred to 30 in. 

barom. ) 

Average superheat, degrees F 

Average load on generator, kilowatts 

Steam consumption, pounds per kilowatt-hour 

Net correction, per cent 

Corrected steam consumption, pounds per 

kilowatt-hour 



75o 
177-5 

27-3 
95-7 
9830-5 
i5-i5t 



Corrected 
to 



179 

28.5 
IOO 



14-57 



Correction, 
Per Cent.* 



-O.15 

-O.30 



* The following corrections were given by the manufacturers and accepted by the purchaser 
as representative of this type and size of turbine: 

Pressure correction o.i per cent, for I pound. 

Vacuum correction 2.8 per cent, for 1 inch. 

Superheat correction 7.0 per cent, for 100 degrees F. 

The percentage correction for pressure in this test becomes then 0.1 (179.0 — 177.5) or 0.15 
per cent. Similarly for vacuum, correction is 2.8 (28.5 — 27.3) or 3.36 per cent.; and also for super- 
heat, correction is 7.0 (100 — 97.5) -5- 100 or 0.30 per cent. These values are tabulated in the 
last column of the table. 

t This is 75 per cent, better than the manufacturer's guarantee. 



9000-KILOWATT CURTIS TURBINE, FISK STREET STATION, COMMON- 
WEALTH ELECTRIC COMPANY, CHICAGO. 



Duration of test 

Speed, revolutions per minute 

Average steam pressure, pounds gauge 

Average vacuum, inches (referred to 30 in. 

barom.) 

Average superheat, degrees F 

Average load on generator, kilowatts 

Steam consumption, pounds per kilowatt-hour 

Net correction, per cent 

Corrected steam consumption, pounds per 

kilowatt-hour 



75o 

179 

29-55 
116 
8070 
13.0 



Corrected 
to 



179 

28.5 
IOO 



14.40 



Correction, 
Per Cent.* 



+9-45 
+9-28 



+ 10.73 



The following percentage corrections were used: 

Superheat correction 8 per cent, for ioo° F. 

Vacuum correction 9 per cent, for 1 inch from curve in Fig. 78. 

Pressure correction not given. 



C. E. Bulletin, 
No. 4531- 



170 THE STEAM TURBINE 

These results show a difference of only .20 pound in the cor- 
rected steam consumption, so that for exactly the same con- 
ditions these two machines would probably give approximately 
the same economy. Each turbine is doubtless best for the 
special conditions for which it was designed. 

These results are equivalent to respectively 9.58 pounds and 9.72 
pounds per indicated horsepower, assuming 97 per cent as the effi- 
ciency of the generator and 91 per cent as the mechanical efficiency 
of a large Corliss engine according to figures given by Stott.* 

From experience with other similar turbines it seems as if the 
vacuum corrections given are too low for each turbine. The 
correction for the Curtis turbine was obtained from the curve in 
Fig. 78 as given between 27 and 28 inches, while it was used 
between 28.5 and 29.5 inches, where the curve of steam consump- 
tion most likely slopes somewhat as shown by the dotted curve in 
the figure, which was derived from the percentage change of 
the theoretical steam consumption calculated from the available 
energy. The correction of 2.7 per cent, per inch of vacuum 
for the Westinghouse-Parsons turbine is probably too low also, 
although the percentage correction would not be nearly as 
large as for the Curtis. If both of these corrections are too 
low, the effect of increasing them would be to increase the cor- 
rected steam consumption of the Curtis turbine and reduce that 
of the Westinghouse-Parsons. 

Large sizes of steam turbines are also made by the Allis- 
Chalmers Company, but sufficient data are not given with pub- 
lished tests to make a comparison here. 

Tests of a 5000-kilowatt Curtis and a 7500-kilowatt Westing- 
house-Parsons turbine are also recorded here for comparison. 
The two tests are corrected to the assumed standard conditions 
of 173.7 pounds gauge pressure, 28 inches vacuum, and o° F. 
superheat. For the test of the Curtis machine the same per- 
centage corrections were used as for the 9000-kilowatt turbine; 

* Electric Journal, July, 1907. It is stated also in this article that the vacuum 
correction of a Westinghouse-Parsons turbine is 3.5 per cent, per inch between 28 
and 28.5 inches. Jude states that the vacuum correction for Parsons turbines 
(English) is five to six per cent. 



METHOD FOR CORRECTING STEAM TURBINE TESTS 



171 



and for the test of the Westinghouse turbine the vacuum correc- 
tion is that given in the footnote at the bottom of page 170 (3.5 
per cent, per inch), while the other percentage corrections are the 
same as in the preceding test of a similar machine. The West- 
inghouse turbine was operated with wet steam. In a test of a 
reciprocating engine the equivalent economy with dry steam is 
calculated by merely subtracting the percentage of moisture, but 
in a turbine test the correction is generally stated as being a little 
more than twice the percentage of moisture.* In other words, in a 
turbine test the moisture must be subtracted twice. The reason 
for this difference in the methods of correcting water rates of 
engines and turbines is the very large increase in the disk and 
blade rotation losses in wet steam (cf. Fig. 69). 



5000-KILOWATT FIVE-STAGE CURTIS TURBINE, L STREET STATION, 
BOSTON EDISON COMPANY. Tested Jan. 29, 1907. 







Corrected 
to 


Correction, 
per Cent. 


Duration of test, houjs 










720 

173-7 
28.8 

142 

5i95 
I3-5 2 






Average steam pressure, pounds gauge 


173-7 

28 

O 


O 


Average vacuum, in. (referred to 30 in. barom.) . . . 
Average superheat, degrees F 


+ 6.40 
+ II.36 


Average load on generators, kilowatts 

Steam consumption, pounds per kilowatt-hour. . . 
Net correction 








+ I7-76 


Corrected steam consumption, pounds per kilo- 




15-92 









7500-KTLOWATT WESTINGHOUSE-P ARSONS TURBINE (SINGLE FLOW 
TYPE), INTERBOROUGH RAPID TRANSIT COMPANY, NEW YORK. 



Corrected 
to 



Correction, 
per Cent. 



Duration of test, hours 

Speed, revolutions per minute 

Average steam pressure, pounds gauge 

Average vacuum, in. (referred to 30 in. barom.) . . 

Average moisture, per cent 

Average load on generator, kilowatts 

Steam consumption, pounds per kilowatt-hour 
(wet) . . . 

Net correction 

Corrected steam consumption, pounds per kilo- 
watt-hour • 



149.7 

27. 70 

7i35 
17.79 



173-7 

28 

o 



16. 



2.4 

1.05 

6.0 



■9-45 



* This correction for moisture has been determined by experiment. 



172 



THE STEAM TURBINE 



It is stated that the steam consumption of the Interborough 
Company's turbine is 15.87 pounds at full load and 15.54 pounds 
at 9000 kilowatts when the overload valve opens. The gen- 
erator connected to this turbine is rated at only 5500 kilowatts. 
With a generator more nearly the rating of the turbine it is 
probable still better results would be secured. 

Corrected tests of a 2000-kilowatt Curtis and a 1000-kilowatt 
Westinghouse-Parsons turbine-generator are also given here. 
Assumed standard conditions and corrections are taken the same 
as in the two tests preceding, except that the Westinghouse test 
is corrected to the steam pressure of the Curtis test. 

2000-KILOWATT CURTIS TURBINE, COMMONWEALTH ELECTRIC COMPANY 
CHICAGO. Tested May, 1905, by Sargent & Lundy. 







Corrected 
to 


Correction, 
per Cent. 




1.25 
900 
166.3 

28.5 , 
207 
2024 
15.02 






Speed, revolutions per minute 






Average steam pressure, pounds gauge 


166.3 

28 






Average vacuum, in. (referred to 30 in. barom.) . . 


+ 4-0 
+ I6-53 


Average load on generators, kilowatts 


Steam consumption, pounds per kilowatt-hour . . 
Net correction 








+ 20.53 


Corrected steam consumption, pounds per kilo- 
watt-hour 




18.10 









1000-KILOWATT WESTINGHOUSE-P ARSONS TURBINE 

1907, by S. Gilliard. 


Tested 


September, 






Corrected 
to 


Correction, 
per Cent. 




1 
1800 
147.6 
27.02 
•75 
1503-5 

i°55 
13.61 

19-35 












Average steam pressure, pounds gauge 


166.3 

28 



-1.87 
-3-40 
-I.50 


Average vacuum, in. (referred to 30 in. barom.).. . 


Average load on water brake, horsepower 

Equivalent average load in kilowatts (generator 






Steam consumption, pounds per brake horse- 
power-hour (wet) 






Steam consumption, pounds per equivalent kilo- 
watt-hour (wet) 






Net correction 




-6.77 


Corrected steam consumption, pounds per kilo- 
watt-hour 




18.04 



METHOD FOR CORRECTING STEAM TURBINE TESTS 173 



Curves in Fig. 79 are given to compare the steam consump- 
tion of a standard 5000-kilowatt turbine-generator and a 4-cyl- 
inder compound 5000-kilowatt reciprocating steam engine of 



• ^ § 




" * & 1 


"2! <s-' i& *0 

■%g- -i§- Si" 


S^ .^ \ J&* 


S ^§ M k 


N v V t 


X $ 


v \- 


\ \ 


> 3 it 


S L ■'-'-■ 


\ t - 


Vl 


t4 


- 44 


H t 


t t 




- l 


5 t 


■ J/ t 


- '■- - 2 -T 


t 


r 

y - • 


/ 




1 





l 



SP . 
3? 3 

to ,G 

■ T <u 

^ 8 

a o 

•£ »T 

•r *v 

J I 

C !>> 

W M 



3 5 



B. 



» s $ 



$ 5 K 



u 

o ^ 



i- o 



! § 

m a 
8^ 

O D 



i8- 
s H 



the type used by the Interurban and Metropolitan Companies 
of New York, assuming both units operating under the same 
conditions. These curves illustrate the good overload economy 



174 



THE STEAM TURBINE 



of the turbine, showing that at 50 per cent, overload the engine 
designed for equal work in the cylinders requires for the same 
output 43 per cent, more steam than the turbine. 

These results are particularly interesting because the peak 
capacity of a station with a given equipment of boilers and 
auxiliaries is increased in proportion to the reduction of steam 
consumption at overloads. 

For a given investment the turbine gives a much larger range 



.50 
<Z .45 

I M 

8 -35 
^ 30 



C 26 
7 25 

& 24 

8.28 

13 22 

ft 

I 21 

a 
S 20 

m 18 



\ 


N. 










I 








■500 




^v 


X 








s%&- 


Out'put 












^ 










s! 




3 












v; 


^> 






Ph! 
PS; 




O 




















J! 




W 


309- 


















"^^V 






















£! 
























m | 
























£j 
















V 


1 
























1 


/£ 


W 






c 

n 














x& 








u 
















1 










a 










/ 






V 


1 






1 






















/ 








Steam per E 
9,750 Lbs 


tour 


















1 1 1 

















)0 - 1200 1600 
Speed -Ti.P.M. 



2000 2400 



FlG. 80. Torque, Speed Output, and Efficiency Curves of a Typical 
500-Kilowatt Steam Turbine. 

of load and, moreover, affords the means by which the peak 
capacity of existing stations can be greatly increased. 

The speed output curve (Fig. 80) is very useful to engineers to 
determine if a turbine is running at its best speed. If the cor- 
responding curves of steam consumption per kilowatt output 
(usually called water rate per kilowatt) and efficiency are calcu- 
lated according to the form on page 364, a great deal of informa- 
tion is obtained about the operation and economy of a turbine. 
The torque line in Fig. 80 is always drawn straight, just as a 



METHOD FOR CORRECTING STEAM TURBINE TESTS 175 

Willans flow line. A curve of total steam consumption is usually 
a straight line for the normal operating limits of a turbine, but 
usually becomes curved when a by-pass valve opens on overload, 
or when the turbine is over its capacity so that the pressures 
are not normal in the stages. 

The torque line shows why a turbine engine is not adaptable to 
automobiles. The starting torque of a small commercial tur- 
bine is not large, so that starting would be difficult with a small 
wheel, and reversing and speed reduction would be as difficult 
as with a gasoline engine. The reciprocating steam engine as 
well as the gasoline engine has, therefore, advantages over the 
steam turbine for this service. 



CHAPTER VII. 
COMMERCIAL TYPES. 

In some respects the order in which the commercial types of 
steam turbines are discussed on the following pages is somewhat 
arbitrary; but, essentially, it is in the order of relative simplicity. 
De Laval and Parsons, of the modern designers, were first in 
the field. They were in fact pioneers in the development of 
commercial steam turbines, and other designers have followed 
more or less in their steps. The reasons for giving precedence 
to the types which they developed are therefore obvious, and no 
other explanation is needed. 

Because of its greater simplicity the commercial De Laval is 
first discussed, and is followed with descriptions of the various 
forms of the Parsons turbine and the more recent types. 

DE LAVAL STEAM TURBINE. 

Rational engineering development is nowhere better exempli- 
fied than in the successful performance of the De Laval steam 
turbine. In nearly every respect, even to details, some of these 
designs are still practically the same as the turbines designed 
under the personal direction of De Laval. 

The essential elements of the original De Laval turbine are: 
(i) the nozzles in which the steam expands; (2) a wheel or 
disk with suitable blades; (3) a slender shaft on which the 
wheel is mounted; and (4) a set of reducing gears to change 
the high speed of the turbine shaft to a lower speed adaptable 
for driving machinery. 

Drawings of a small De Laval turbine are shown in Fig. 82. 
The turbine wheel, W, is supported upon the flexible shaft 
between the bearing, Z, provided with a spherical seat, and the 
gland or stuffing-box, P. Teeth are cut into the metal of the 
turbine shaft to make the pinions on each side of K fit the gear 

176 



COMMERCIAL TYPES 



177 



wheels A and B, from which the power is transmitted. The 
design shown here is intended for driving two electric generators 
which are direct-connected by means of the couplings shown at 
the left in the figure. 

De Laval turbine-generator sets of from 50 horsepower up- 
wards are supplied with two gear wheels, two power shafts, and 
two dynamos for each turbine wheel, while the smaller sizes 




Fig. 82. Section of a De Laval Single Stage Turbine with Two Power Shafts 
(on Gear Wheels A and B). 



have gear arrangements for a single generator. Because of the 
higher speed at which the small sizes operate (see page 180), 
making the pressure on the gear teeth considerably smaller than 
with the larger sizes, more power can be transmitted with a single 
set of gears. The large size of the gear wheels compared with 
the turbine is a noticeable feature of these turbines. 

Turbine Wheel. On account of the very high speeds at which 
these turbines operate, the wheels or disks require very careful 
designing. In the small and medium sizes, a wheel similar to 



i 7 8 



THE STEAM TURBINE 



the drawing in Fig. 83 is used. When this design is used, the 
hub of the wheel is bored out and a thin steel bushing is drawn 
into it by means of a nut shown in the figure at the right-hand end. 
Before this bushing is put into the wheel, it is forced on the shaft 




Fig. 83 



De Laval Turbine Wheel with a Hole at the Center and Details of the 
Blades. 



and pinned in place as shown. The wheel can be removed from 

the shaft by taking off the nut and drawing it from the bushing. 

The strength of a disk, or a wheel of a disk type, in which there 

is a hole at the center is at best not more than half as strong as one 




Fig. 84. De Laval Turbine Wheel without a Hole at the Center. 

without a hole.* On this account in the larger sizes of De Laval 
turbines it has been found necessary to use the design shown 
in Fig. 84. In this arrangement a solid disk is permitted. 
The hub is recessed at each end, and the flexible shaft is made 

* An explanation of this remarkable phenomenon is given on page 425. 



COMMERCIAL TYPES 179 

with enlarged flanged ends which fit into the recesses and are 
bolted solidly in place. The recesses and flanges are machined 
with a four per cent, taper in order that the parts maybe accurately 
centered and fitted. 

This form of wheel disk with the section increasing from the 
rim towards the hub is arrived at by proportioning it to have 
equal unit stresses throughout. But this condition does not 
hold true at the rim, where just below the blades annular 
grooves are turned on each side. Weakening of the wheel at 
the rim is a very good method of providing for abnormal stresses 
that result in case of a failure of the governor to control the speed. 
The purpose in making these grooves is to have the wheel burst 
at this reduced section where the stresses per unit of area are 
about 50 per cent, larger than at any other part of the wheel, 
rather than near the center where the damage from failure would 
be so much greater. At normal speed the factor of safety, at 
this smallest section, is about five, and since the unit stresses 
vary as the square of the speed,* the wheel will fail at this place 
at a little more than twice the rated speed. As these wheels are 
constructed, no great damage to the turbine will result, therefore, 
from the failure of the wheel rim. It has been shown by actual 
experiments with such wheels that when failure occurs, the rim 
holding the blades is broken up into very small pieces which 
will not damage the wheel case. It is stated, however, that 
wheels without this reduced section, when tested to failure, have 
been broken up into two or three large pieces by bursting through 
the center, and these pieces have been driven through an experi- 
mental wheel casing made of two-inch steel castings. 

There is also another consideration that is especially interesting 
to engineers. When a portion of the rim breaks off the wheel 
becomes unbalanced, and as the clearance between the heavy 
hub of the wheel and the safety bearings in the surrounding 
casing is very small, as can be seen in Fig. 82, the flexibility of the 
WV 2 

* Centrifugal force = (see page 406) and is therefore proportional to the 

square of velocity (speed). The factor of safety at other sections of a De Laval 
wheel is about eight. 



i8o 



THE STEAM TURBINE 



shaft will permit the hub of the wheel to come into contact with 
the circular openings in the casing into which it extends. The 
friction of these surfaces will act as a brake and assists in bringing 
the wheel to rest. And this is easily accomplished, because with 
the blades removed the steam no longer acts to rotate the wheel. 
The diameters of the wheels are relatively small, as can be seen 
from the following table: 



Horsepower 

Revolutions per minute 

Diameter to center of blades, inches. 
Blade speeds, feet per second 



5 
30,000 

3-94 
523 



30 
20,000 
8.86 

785 



100 
13,000 
19.68 

ii34 



300 
10,000 
29 .92 

1310 



Wheels for De Laval turbines are usually made of a special 
forged nickel steel said to be rather high in carbon. 

Nozzles. Fig. 85 is a typical illustration of a 20-kilowatt 
De Laval turbine-generator and gives a general idea of how the 
nozzles which direct the steam against the blades are arranged 
around the periphery of the turbine wheel. 'They are attached 
to the turbine mechanically by being fitted into the circumfer- 
ence of the steel casting which serves as the casing for the wheel. 
The number of nozzles varies according to the size of the turbine. 
The nozzles are provided with hand valves, which can be seen 
in the figure, by which they can be closed when the turbine is 
running at light loads. In this way some of the nozzles are " cut 
out" and a relatively high efficiency is obtained at light loads. 
In this particular case, about half of the openings in the casing 
for nozzles are closed by plugs; but by removing these plugs 
and inserting nozzles instead, the capacity of the turbine would 
be greatly increased. 

The nozzles are the only parts of a De Laval turbine that 
are changed to make it suitable for any particular pressure, 
degree of superheat, or vacuum. The ratio of the admission 
(usually boiler) pressure to the exhaust pressure is the most im- 
portant factor influencing the design of a nozzle. Briefly stated 
this ratio of pressures determines the areas of the cross-section 
of the nozzle at the throat and at the mouth, and therefore its 
divergence or taper. 



COMMERCIAL TYPES 



181 



For the same output more steam is required at a low pressure 
than at a higher pressure. De Laval turbines are readily adjusted 
for a change of boiler pressure by adding more nozzles if they are 




needed. Sometimes turbines are fitted with two sets of nozzles, 
one suitable for condensing and the other for non-condensing 
operation. 



102 THE STEAM TURBINE 

Reamers are used to produce the required taper on the inside 
of these nozzles. In the works at Trenton over 600 reamers are 
kept in the tool room. The taper of the nozzle ranges from six 
to twelve degrees, and the clearance between the mouth of the 
nozzle and the blades (axial clearance) is about an eighth of an 
inch. 

Blades. De Laval blades are made of drop-forged steel and 
have bulb shanks which are fitted into suitable slots in the wheel, 
shown in Fig. 83, which are milled across the rim and then drilled. 
The blades are lightly calked to secure them in place. At the 
upper ends of the blades they are provided with " extensions" 
which are designed to make adjoining blades fit closely and thus 
form a continuous ring over the blades at the periphery of the 
wheel. Details of these blades are shown more clearly in 
Fig. 64. 

Shaft. Small De Laval turbines have two important features 
distinguishing them from all other types. The first is the long 
diverging nozzle with the hand wheel control already mentioned; 
and the second is the slender flexible shaft * of the turbine. A 
wheel revolving at a very high speed tends to rotate about its 
center of gravity. If it is mounted on a stiff, unyielding shaft, of 
which the axis does not pass through the center of gravity of the 
wheel, this tendency causes violent vibrations of the wheel and 
shaft due to the very large centrifugal forces. It is stated that a 
weight of one ounce attached at the circumference of the wheel of 
a 300-horsepower De Laval turbine will produce an unbalanced 
centrifugal force of nearly 2000 pounds. It is mechanically 
difficult and almost impossible to construct a wheel so perfectly 
balanced that its center of gravity will exactly coincide with 
the geometric center of the shaft on which it is mounted. 
De Laval, therefore, devised a long, slender shaft which, as the 
speed of the wheel increases, yields somewhat and allows the 
latter to assume its own position of rotation about its center of 
gravity. 

* The diameter of the shaft of a ioo-horsepower De Laval turbine is i inch and 
of a 300-horsepower turbine is about i^ inches. 



COMMERCIAL TYPES 183 

The wheel is not mounted midway between the bearings but 
considerably nearer the spherical seated bearing Z, Fig. 82, at 
the governor end. Now when the wheel is started up from rest, 
if its center of gravity is not precisely in the axis of the shaft, it 
will bend, and the plane of revolution of the wheel is then no 
longer perpendicular to the axis of rotation. When, however, a 
sufficiently high speed is reached, so that gyroscopic action is 
great enough to pull this plane back to a position perpendicular 
to the axis of rotation, a "node" is formed at the center of the 
hub and rotation will then take place about the center of gravity 
of the system. The speed at which the amplitude of vibration 
is greatest is called critical.* 

Bearings. Typical bearings of De Laval turbines are illus- 
trated in the section drawings in Fig. 82. At the right-hand or 
"governor'' end there is a spherical seated bearing (Z). A 
design of this kind is used for the purpose, primarily, of giving 
greater flexibility to the shaft and to take the small end thrust 
exerted on the wheel by the steam issuing from the nozzles at a 
very high velocity. In single wheel turbines of the De Laval 
type this pressure or thrust is, however, very slight, as the steam 
is expanded to the exhaust pressure before it leaves the nozzles. 
It is obvious, therefore, that the wheel rotates in steam of very 
nearly the same pressure on both of its sides. Such a design 
has also the advantage of being self-aligning. A helical spring 
shown in the same figure holds the spherical bearing against its 
seat in the turbine casing. On the other side of the turbine 
wheel the shaft passes through a loose-fitting bearing, P, serving 
primarily as a gland or stuffing-box to prevent the leakage of 
steam from the casing. The shaft does not pass through the 
casing on the right-hand side, so that no precautions are necessary 
to prevent leakage of steam on that side. At each side of the 
pinions of the reduction gearing, the turbine shaft is supported 
on plain white-metal (Babbitt) bearings C and CC. The sur- 

* " Critical speed " is the name given to that speed of a wheel at which it tends 
to rotate about its own center of gravity. In the De Laval turbines a critical 
speed occurs at about i to | of the normal running speed. 



184 



THE STEAM TURBINE 



face speed in these bearings is usually designed to be about 
70 feet per second. 




B 



Oh 

b 

PL, 



C tf 
I ^ 

P 5 

d 
IS 



Speed-reduction Gears. On account of the high speed of the 
turbine shaft, reduction gears are required to bring the speed 
within practicable limits for utilizing the power. The reduction 



COMMERCIAL TYPES 185 

is usually about ten to one, and is accomplished by means of 
small pinions on the turbine shaft meshing with steel helical 
gear wheels. The teeth of the pinions are very small and are 
cut directly into an enlarged section of the flexible shaft.* The 
teeth for this gearing are cut spirally at an angle of 45 degrees. 
As indicated in Fig. 86 the teeth on one side are cut on a right- 
hand and on the other side on a left-hand spiral. This method 
effectually prevents any movement of the shaft in the direction 
of the axis and balances the thrust of the gears. Previous to 
the time when De Laval demonstrated that gears could be oper- 
ated at a linear velocity of more than 100 feet per second, the 
high speeds which he introduced were not considered practically 
possible. His success at these high speeds was due largely to 
the fine pitch | and spiral angle of the teeth. It is thus possible 
to bring a large number of teeth into mesh at the same time, 
so that the working pressure on each tooth is made very small 
and abrasion is reduced to a minimum. 

The reduction gears are enclosed in a casing entirely separate 
from that around the turbine wheel. This casing prevents dust 
and grit from getting into the gears and avoids accidents from 
persons or objects falling upon them. With careful attention 
these gears sometimes run for several years without visible wear. 
Formerly the gear wheels were made of bronze, but experience 
showed that the teeth became crystallized after a few years of 
operation, and pieces of the teeth which were sometimes broken 
off, were liable to injure other teeth. Such gears should always 
be supplied with a little oil for lubrication. 

This speed-reduction gearing introduces two important dis- 
advantages: first, the friction loss is considerable; and second, 
the construction is necessarily expensive. The friction loss, 
obviously, will depend largely on the quality of workman- 
ship. It is stated that this loss in the gears is about 5 per 

* The pinions are said to be made of .60 to .70 carbon steel, and the teeth of the 
larger gear wheels are cut in .20 carbon steel of a grade similar to that used for 
locomotive wheel tires. 

t The pitch of the gears varies from .15 inch in the smallest, to .26 inch in the 
largest sizes. 



1 86 THE STEAM TURBINE 

cent.* of the power transmitted when they are in good condi- 
tion, and sometimes as much as 10 per cent, in moderately 
worn gears. 

After a few years of service it is usually found that the steam 
consumption of a De Laval turbine is slightly greater than when 
it was new. This poorer economy is probably due to the in- 
creased loss in the gears from wear as well as to the wearing 
away of the blades on the turbine wheel, which by changing the 
shape of the blades causes a loss of efficiency. 

Governor. Types of De Laval governors are shown in Figs. 
162, 163, and 174b, pages 276 and 301, where methods of gov- 
erning are discussed. The valve arrangement controlled by the 
governor is a plain throttling type. 

DE LAVAL MULTI-STAGE TURBINES. 

For a number of years the De Laval Company confined its 
efforts to the construction of single stage turbines described in 
the preceding sections. In the last few years there has been a 
growing "demand for larger sizes and multi-stage De Laval tur- 
bines have been designed to meet this demand. Single stage 
turbines are obviously of limited capacity. The maximum out- 
put of a thirty-inch De Laval turbine with a single wheel is 
about 500 kilowatts. In order to increase the capacity with- 
out increasing the diameter of the turbine wheels beyond the 
practicable limits of construction, it is necessary to resort to the 
use of multiple stages. By using a sufficient number of stages 
the speed of the turbine can be reduced to any desired point, 
the designer given practically a free choice as to the proper 
relations of wheel diameters, capacity, speed, length of blades, 
and other variables. After the most suitable speed for best 

* Regarding these losses the results of experimenters differ a great deal. Lewicki 
found the gearing and bearing loss in a 30-horsepower De Laval turbine-generator 
to be 7.5 per cent, of the full load output. Delaporte states that the gearing losses 
of a 200-horsepower De Laval turbine are about 1 per cent, when new; and he states 
also that in his opinion the combined gearing and bearing friction losses of a 300- 
horsepower De Laval turbine should be taken roughly at about 3 per cent. 



COMMERCIAL TYPES 3:87 

efficiency has been selected, any further reduction in the speed 
of the turbine as may be necessary to adapt it to driving a par- 
ticular machine may be accomplished by means of a suitable 
reduction gear. For example, De Laval multi-stage turbines, 
when provided with reduction gears, are suitable for direct 
connection to standard speed, direct current generators ena- 
bling the user to escape all disadvantages inherent in high speed 
direct current generators, the normal designs of which are ordi- 
narily changed in many respects when they are to be adapted 
to run at the high speed of turbines. Likewise, the speeds most 
suitable for centrifugal pumps and blowers are, for many serv- 
ices, lower than the speed at which turbines should be oper- 
ated, and these, also, may be driven through the medium of 
gears. The use of gears permits the use of turbines for belt or 
rope driving with pulleys and sheaves of ordinary dimensions 
and running at standard " rim " velocities. Further, a De 
Laval multi-stage geared turbine can be connected to existing 
shafting by belts to supplement the power delivered by recip- 
rocating engines already in service. 

Turbines of this type, however, are also supplied without 
gearing, for driving machines the speed of which approximates 
that of the turbine, such as large alternators, centrifugal air 
compressors, and high head centrifugal pumps. 

For any given capacity and steam condition there is, for each 
type of turbine, a speed which will give the highest efficiency, 
.and it is a problem for the designer to make the proper arrange- 
ment of blade lengths, blade angles, disk diameters, etc., so as 
to secure the ratio of steam velocity to blade speed that will 
secure the best efficiency. In medium sizes, that is, up to 
about 3000 kilowatts, the multi-stage impulse turbine, in some 
cases fitted with velocity stages in the first pressure stage, is the 
most efficient by quite an appreciable margin. Comparing the 
multi-stage turbine with the reaction turbine in small sizes 
under these conditions, the high pressure blades of the reaction 
turbine become very short in proportion to the leakage path 
through the clearance at the ends of the blades. Due to the 



1 88 



THE STEAM TURBINE 




COMMERCIAL TYPES 



189 



fact that the leakage takes place over the ends of the moving 
blades and under the ends of the stationary blades of reaction 
turbines, it is not feasible to use types of packing which is ap- 
plied so effectively for packing the joints between the shafts 
and diaphragms of multi-stage turbines. 

A sectional drawing of a typical De Laval multi-stage steam 
turbine is shown in Fig. 87. The rotating part of the turbine 
consists of a heavy shaft upon which is mounted a series of disks 
for carrying the blades. The shaft and the blade disks are shown 
in Fig. 88. Each blade disk revolves in an independent cham- 




r~ 



km**** 



Fig. 



Shaft and Blade Disks of De Laval Multi-stage Turbine. 



ber formed between diaphragms held in a cylindrical casing. 
Steam is admitted to the steam chest at the right-hand end of 
the casing and then flows through the nozzles of the first stage 
and discharges upon the blades of the first disk.* Next the 

* In the latest designs of these turbines, there are two blade disks in the first 
stage with a set of intermediate blades, making two velocity stages similar to 
Fig- 39- 



190 



THE STEAM TURBINE 



steam flows through guide vanes in the diaphragm separating 
the first stage from the second and impinges upon the second 
blade disk, and so on through succeeding stages of the turbine. 
The nozzles of the first stage occupy only a portion of the cir- 
cumference, thereby avoiding the difficulties of very short blade 




Fig. 89. Diaphragm (showing nozzles) of a De Laval Multi-stage Turbine. 

lengths which would otherwise be necessary if the admission of 
steam were permitted all around the circumference in this 
stage. 

The blades or " buckets " of all De Laval turbines are drop 



COMMERCIAL TYPES 



191 



forgings and the bulb shanks are accurately machined to fit the 
corresponding recesses in the blade disks. 

With the exception of the nozzles in the first stage, which have 
been described in connection with the single stage De Laval 
turbine, the nozzles of the succeeding stages are formed by the 




Fig. 90. Carbon Packing on Shaft of Turbine. 



accurate locating of adjacent guide vanes made of nickel-bronze, 
accurately formed in dies and finally hammered to give the 
surface a hard polish, which improves the density and strength 
of the metal. They are spaced and located upon the rim of 
the diaphragm by pins and are held in place by a solid steel 
band shrunk over their tips. Two pins are used for each vane 



192 



THE STEAM TURBINE 



to determine its proper angle, and therefore, in connection with 
the shape of the vanes, to fix the contour and the cross-sectional 
area of the nozzles formed between the successive vanes. 

The cast-iron disks used for the diaphragms are perforated 
at the center and are fitted with removable labyrinth packings 
in order to minimize the leakage of the steam from stage to stage 
between the diaphragms and the cylindrical wheel hubs. A 
complete diaphragm of a De Laval multi-stage steam turbine is 
shown in Fig. 89. 















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80 100 120 140 
Superheat -Degs. Fahr. 



160 180 200 



Fig. 91a. Percentage Curve for Correcting De Laval Turbine Tests for 

Superheat. 



The carbon packing used at the low-pressure end of De Laval 
multi-stage turbines is shown in Fig. 90. Provision is made for 
introducing live steam at reduced pressure between the second 
and third rings as shown so that any leakage into the turbine 
casing will be steam from this source and not air from outside 
the casing. 

Superheat, Vacuum, and Economy Curves. Fig. 91a shows 
by percentages the effect of superheat on the steam consump- 



COMMERCIAL TYPES 



193 



tion. For low values of superheat the gain for a De Laval tur- 
bine is much greater than for larger amounts of superheat. 
Such curves on a percentage basis are sometimes very service- 
able to show striking variations clearly. Fig. 91b is a similar 
percentage curve to show how the vacuum influences the steam 
consumption. With a high vacuum the improvement in econ- 
omy is much more marked than at low values. Fig. 91c shows 
approximately the steam consumption for small sizes of De Laval 



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Vacuum iu Inches of Mercury 



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28 



Fig. 91b. Percentage Curve for Correcting De Laval Turbine Tests for 

Vacuum. 



turbine-generators operating non-condensing or with 28 inches 
vacuum at 165 pounds per square inch absolute pressure, and o 
degrees F. superheat. 

It is stated the half load steam consumption of a single stage 
De Laval turbine is 12 per cent, greater than at full load, and 
that at quarter load it is 25 per cent, more than that at full load. 
For such good performance at light loads it is necessary to 
operate the turbine with no more valves open than are needed. 



194 



THE STEAM TURBINE 



Because the valves must be operated by hand such good economy 
could probably not be obtained with a rapidly fluctuating load. 







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100 120 140 160 180 200 220 240 
Bated Full Load Output — JKw. 

Fig. 91c. Approximate Steam Consumption of Small Sizes of De Laval Turbrne- 

Generators. Dry Saturated Steam at 165 Pounds per Square Inch Absolute 

Pressure. 

Turbine Losses. The following table shows how the losses in 
a single stage De Laval 200-kilowatt turbine-generator have 
been divided up by Stevens and Hobart: 

Nozzle losses 

Radiation losses and leakage 

Rotation losses due to the turbine wheel revolving in steam 

Losses due to the steam traveling over the blades 

Bearing friction losses 

Losses in speed-reduction gearing 

Generator losses 

Losses due to residual kinetic energy in the steam passing 

to the condenser 8 

Electrical output 59 

Total .. . 100 



[2 per cent. 

j Ct (« 

4 " " 

9 « « 

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2 " " 

4 « « 



COMMERCIAL TYPES 195 



PARSONS TURBINE 



The Parsons type of steam turbine differs from that commonly 
known as De Laval's principally in the substitution of stationary 
blades in the place of nozzles. These stationary blades are so 
shaped as to direct the steam upon the moving blades just as 
nozzles would. In turbines of this type a large number of rows 
of moving blades are employed, which are attached to the cylin- 
drical surface of a revolving drum, called a rotor. 

There is also another difference which, from a theoretical view- 
point, makes a Parsons turbine entirely different from other types. 
All the impulse turbines, of which the De Laval is a good example, 
make very little, if any, provision for the expansion of the steam 
in the moving blades, while the Parsons type is designed to give 
approximately as much expansion of the steam in the moving 
as in the stationary or "guide" blades. In turbines of this 
type each set of one row of moving and one row of stationary 
blades is called, technically, a stage. 

Compared with the De Laval turbine in which the blades of 
a single wheel revolve in a medium of uniformly low density 
with the pressure very nearly the same on both sides of the wheel, 
most of the blades of a Parsons turbine revolve in steam of high 
density. Blades at the admission end revolve in steam at very 
nearly the boiler pressure, and only those at the low-pressure end 
are in steam of low density. 

In the Parsons turbine there is a considerable drop in pressure 
in every row of blades, and consequently a difference in pressure 
between the two sides of every row, which produces a leakage 
of steam over the edges of the blades, increasing with the amount 
of radial clearance between the stationary and moving parts. 

Because of the large number of blades, this leakage of steam 
is a factor which on account of its magnitude must receive most 
careful attention and investigation by designers. It is a matter 
of the greatest importance, therefore, in designing turbines of 
the Parsons type to make radial clearances as small as possible, 
consistent with proper allowances for the expansion due to un- 



196 



THE STEAM TURBINE 



equal heating of the parts,* which in a turbine with a large 
number of stages is a very important consideration. Fig. 91 d is 

a section of a typical Par- 
sons rotor and casing show- 
ing by arrows the leakage 
spaces for steam through 
the radial blade clearances 
a and b. 

A section of one of the 
early Parsons turbines is 
illustrated in Fig. 92. The 
turbine rotor consists of a 
long drum of three different 
sections supported on the 
two bearings — one at each 
end. The moving blades 
are mounted on the cir- 
cumference of this drum 
and the stationary blades 




^^^^^f^^^^^^^M 




mm 



Fig. 9id. Section of a Typical Parsons Rotor are fitted in similar rings 
and Casing Showing the Radial Blade to the fa^fe Q f tne tur b m e 
Clearances. 

casing. 

The annular space I is the steam chest which receives high- 
pressure steam. The steam passes through the alternate rows 
of moving and stationary blades of the first section of the rotor, 
through a second annular space to the blades of the second sec- 
tion which discharge into a still larger annular space, from which 
it passes through the blades of the last section of the rotor to 
the exhaust E. At the second and third annular spaces, where 
the diameter of the drum is increased, an unbalanced pressure 
or thrust toward the right is produced by the pressure of the 

* Aside from the question of radial clearance, all other points affecting the design 
are of minor importance as regards economical and satisfactory operation. The 
most successful design of a Parsons type is the one which operates successfully 
with the smallest radial clearances. Unequal expansion of the different parts of 
the casing and drum introduces factors which are very difficult to estimate. If the 
blades are made of different materials trom the drum, at some temperatures they 
are likely to be loose. 



COMMERCIAL TYPES 




ft 

.1 



198 THE STEAM TURBINE 

steam; and this thrust is increased by the expansion of the steam 
in the moving blades (see Fig. 34). To balance this axial pres- 
sure, three balance pistons are provided at the left-hand end 
of the casing — one is intended for each section of the rotor. 
The smallest is made just large enough to equilibrate the thrust 
due to the blades of the first section; the intermediate piston 
balances the thrust on the second annular area and that due to 
the blades of the second section; and the largest piston equili- 
brates the pressure on the third annular area and the thrust in 
the third section. Steam passages are cored out in the casing, 
as shown in the figure, to make each balance piston communi- 
cate with its corresponding section of the rotor, so that the pres- 
sure in the section is always the same as that acting on the 
corresponding balance piston. In some designs these cored-out 
passages are replaced by pipes on the outside of the casing. 
Small annular grooves are usually cut in the balance pistons to 
join with similar annular projections in the casing. This con- 
struction, called a labyrinth packing, makes a devious and ob- 
structed steam path * so as to effectually prevent undue leakage 
of steam around the balance pistons. 

The position of the moving blades with respect to the station- 
ary blades (axial clearance) is usually adjusted by means of a 
thrust or adjustment bearing T at the extreme left-hand end of 
the turbine. It consists of a number of rings or collars turned 
in the steel shaft into which corresponding brass rings in the 
adjustment bearing are fitted. The upper and lower halves of 
this bearing are adjustable and are moved by the screws shown 
in the figure. If the lower half of the bearing is set so that the 
collars on the shaft are in contact on their left side, the upper 
half would have the collars in contact on the right side. By 
this means, when the bearing is once set, the rotor cannot move 
an appreciable distance either to the right or to the left. A 
typical adjustment bearing is shown more clearly at the right 
in Fig. 107. In this design the upper and lower halves are 

* The labyrinth packing produces a subdivision of the total drop in pressure 
between the right side of the small drum and the left side of the large drum that we 
have instead of a single pressure drop a large number of small pressure drops (due to 
a whole series of resistances) and the leakage of steam is reduced accordingly. 



COMMERCIAL TYPES 



199 



moved by micrometer screws, so that the axial position of the 
rotor is indicated at all times by the dials on these adjusting 
screws. 

In Fig. 92 a very common method for operating the governor 
of steam turbines is illustrated. A worm gear on the main tur- 
bine shaft engages with a gear wheel which by means of other 
gears rotates the governor shaft. 

Detailed Description of Parsons Turbine. Fig. 92 shows a 
section of a Parsons turbine of the stationary or " land " type. 
It consists of a cylindrical cast-iron casing CC and is made in 
halves, the upper half being removable for purposes of examina- 
tion without disturbing the rotor. The main casting for the 
casing is carried by a saddle or bed-plate B which is in turn 
fastened to the foundation. The turbine is, however, rigidly 
bolted to only one of the bed-plates, and is free to slide at the 
other, when its length changes due to changes of temperature. 
In the figure the casing will expand and slide toward the left 
and carry the thrust bearing T with it; the shaft expands 
from the thrust block toward the right. Since these two effects 
tend to balance each other, there is very little " absolute " 
motion of the coupling at P. Note that the main casting is in 
the position of a beam resting on supports at each end, and it 
must therefore have considerable rigidity to prevent deflection. 

The steam after passing through the emergency and governor 
valves enters the casing at I. It can then pass completely 
around the rotor by means of the " steam belt," formed in the 
cylinder casing. The steam can thus readily enter the annular 
space occupied by the fixed and moving blades. 

Having traversed the full length of this annular space the 
work of the steam is completed, and it is discharged by way of 
the exhaust end EE, and the exhaust pipe E r to the condenser. 

The balance pistons and the pressure equilibrium passages 
are readily observed. One of the equilibrium passages is formed 
in the main casting to connect the annular chamber at the be- 
ginning of the intermediate section with corresponding balance 
piston. A third equilibrium passage Q is connected to the ex- 
haust main, so that the pressure in the space S is the same as 
that in E. 



200 THE STEAM TURBINE 

The second inlet and steam belt, shown at K, is for use when 
the turbine is overloaded. It is connected by means of a by- 
pass and valve with the inlet I. When this valve is opened, the 
live steam enters the second expansion, or series of blades, with- 
out having previously passed through the first. The turbine 

could then generate enough power 
to cope with, say, a 50 per cent, 
overload, though not with the 
same economy as when working 
for its designed output with the 
by-pass closed. 

The bearings shown at b, b are 
of the " flexible " or elastic sleeve 
pattern, and allow of a slight auto- 
matic adjustment of the center line 
of the rotor. The flexible cou- 
pling for connecting the dynamo 
and turbine shaft is shown at P. 
Packing Glands. In every tur- 
bine, glands or stuffing-boxes must 
be provided where the shaft passes 
through the ends of the casing to 
prevent the escape of steam at the 
high-pressure end and the entrance 
Fig. 93. Propeller of a Water-packed of air at the low-pressure end 
Tu a r bbe° f a Westinghouse - Parsons of condensing turbines. Steam- 
packed glands of various types are 
often provided; but in the Westinghouse-Parsons turbine water- 
packed glands are now generally used. This arrangement con- 
sists of the propeller of a centrifugal pump (Fig. 93) which 
rotates in the water supplied to an annular groove in the casing. 
When the turbine is operating the water is thrown outward by the 
vanes and comp etely fills the space around the periphery of the 
propeller. By this means the leakage of steam or air is effectually 
prevented. As there are no rubbing surfaces in these glands and 
no oil is used, there is no contamination of the exhaust steam. 

Blades. The blades of Westinghouse turbines are secured to 
the rotor by means of slots turned on its periphery, which are 




COMMERCIAL TYPES 



201 



narrower at the top than at the bottom. Into these slots the 
blades which have been cut at the roots to fit, are put singly „ 
Soft metal spacing pieces of the required shape to fill the space 
in the slot between the blades are calked to hold the blades 
firmly by a dovetail construction. This construction is required 
for the attachment of the moving blades to give the necessary sup- 
port against centrifugal forces ; but as the stationary blades, which 
are fixed to the inside of the casing, are not subjected to centrifugal 
forces, the slots for these blades are not usually dovetailed. 




Fig. 95. Blades on the Rotor of a Westinghouse Turbine. 



Blade Lashing and Shroud Rings. It has been found necessary 
to bind the blades together at their ends to make a stronger con- 
struction. In the earlier designs of Parsons turbines the blades were 
usually bound together with wires soldered to their ends. Some- 
times, however, the blades were turned over at their outer ends to 
form flanges which were soldered together into a solid shroud. 

Fig. 95 shows several rows of blades of a Westinghouse turbine. 



202 



THE STEAM TURBINE 



All blades more than two inches long are reenforced by lashing 
with a wire of special section threaded through punched holes 
in the ends of the blades. This method of lashing is illustrated 



SffMFLSS OF THE. 

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Fig. 96. Method of Lashing Westinghouse Blades. 

by Fig. 96. The lashing wire, which is drawn to have a cross- 
section resembling a comma, binds the blades together firmly 
enough to give adequate strength for normal service, yet, unlike 
a very rigid blade construction, it will yield in emergencies without 



COMMERCIAL TYPES 



203 



seriously damaging other parts of the turbine. The blades are 
lashed in sections three feet long. Because of the peculiar shape 
of the section of the lashing wire, it can be calked at the end so 
that a "key" remains in the punched hole to prevent the blade 




ALLIS- CHALMERS CO 

PATENTED , \\ 

Fig. 07. Sankey's Blading for Parsons Turbines. 

from getting out of line. In many respects it is practically as 
effective as a shroud ring. 

A type of blading for Parsons turbines, patented by H. R. 
Sankey in 1903, has been applied with certain modifications in 
the Allis-Chalmers and the Willans turbines. A typical illus- 
tration of this blading is shown in Fig. 97. It is distinguished 



204 THE STEAM TURBINE 

principally from the usual Parsons blading by the attachment of 
a U-shaped shroud ring, B, around both the moving and the 
stationary blades. 

The blades are cut to the required length from bars of copper 
alloy drawn, like wire, to a suitable shape. After the blades 
are cut from the bar, they are formed in machine tools of 
special design, so that at the root they have an angular "dove- 
tail" shape as illustrated in the figure, where the blades are shown 
inserted in a suitable foundation ring, A. After this foundation 
ring is turned to the proper diameter, " dovetail" slots for the 
blades (see Fig. 98) are cut by a special milling machine 
intended for very accurate spacing and inclination so as to give 
the required pitch and angle to the blades. 




Fig. 98. Spacing for Sankey's Blading. 

After the roots of the blades have been inserted in the founda- 
tion rings, which, in cross-section, are also of a dovetail shape, 
the rings are inserted into corresponding grooves in the drums 
of the rotor and in the inside of the casing where they are held in 
place by "key pieces." Each of these "key pieces" after being 
driven into place is upset in an undercut groove which serves as a 
locking device. The dovetail shapes used in this construction 
make the attachment of the blades at their roots very secure. 

The channel-shaped shroud rings are purposely made thin at 
the flanges so that in case of contact between the revolving and 
stationary parts these flanges will be worn off at their edges 
without tearing out or bending the blades. By this method, as 
well as with all other types of shroud ring construction, the strength 
of the blading depends, not on the strength of a single blade, 
but on the total strength of as many blades as are bound together. 
In the Allis-Chalmers turbines all the blades in a semi-circum- 



COMMERCIAL TYPES 



205 



ference are joined by a shroud ring. The blading is thus made 
up in half rings, which are made almost entirely by machinery. 




Each ring can be thoroughly inspected before being placed in the 
turbine and the possible inaccuracies of hand work are likely to 



206 



THE STEAM TURBINE 



be eliminated. Fig. 99 shows the interior of the casing of a 
turbine fitted with shroud rings on the blades. 

If small radial clearances are desired, exceptional precautions 
in designing must be taken to avoid unequal expansions of the 
parts of the rotor, the casing, and the blades, because shroud 
rings in reaction turbines are liable to produce disastrous results 
by " stripping" the blades. Usually in case of accident, however, 
damaged or worn rings can be removed and the turbine continued 
in operation until they can be replaced. 




Fig. 100. A Westinghouse High Speed Flexible Bearing. 

Bearings. In turbines of the Parsons type operating at above 
1800 revolutions per minute, a design of flexible bearing (Fig. 100) 
is used to reduce the vibrations of the shaft by permitting the 
rotor, when passing its critical speed, to revolve about its center 
of gravity instead of its geometric axis. This flexible bearing 
consists of a nest (usually four) of loosely fitting cylindrical 
bronze sleeves between which oil films are maintained by capil- 



COMMERCIAL TYPES 207 

lary attraction.* The clearance between these sleeves is about 
.004 inch. These films of oil have also a cushioning effect in 
absorbing vibrations that occur when bringing the turbine up to 
speed. This flexible bearing accomplishes the same purpose for 
which De Laval used a flexible shaft. In the figure the outer 
casing of the bearing is at the right-hand side and the holder 
for the Babbitt metal lining and the cylindrical sleeves around 
it are shown at the left. 

In larger machines which run at lower speeds, balancing is less 
difficult and single spherical-seated bearings lined with Babbitt 
metal are used. Quadrant liners are provided for either 
type of bearings to accurately adjust the rotor to a central 
position. 

Stages. In this type of turbine low blade speeds are secured 
by using a larger number of stages. Thus in a 400-kilowatt 
" single-flow " Parsons turbine there are 58 stages or 116 rows of 
blades. In such a turbine there are about 30,000 blades. It is 
important to notice why the pressure difference for each row of 
blades gradually decreases from the admission to the exhaust in 
such a turbine. Since there are 58 stages, if the pressure dif- 
ferences were made equal for a total drop in pressure of say from 
175 pounds per square inch to 1 pound per square inch, the drop 
in pressure in each stage would be 3 pounds per square inch. 
But because the steam velocity for a given difference in pressure 
is very many times as great at 1 pound as at 175 pounds, such a 
division is not desirable, and instead the pressure drop is made to 
suit blade speeds that are likely to show best efficiency in the 
various sections. Minimum and maximum velocities at the 
low-pressure end are 500 to 700 feet per second in modern 
designs of these turbines. 

A large Westinghouse-Parsons turbine is shown in Fig. 101, 
with the upper half of the casing removed to show the rotor, 

* Bearing pressure in pounds per square inch times peripheral velocity of the 
shaft in feet per second is generally about 2500. — Proc. Inst. Elec. Engrs., June, 
1905. 



208 



THE STEAM TURBINE 




a 



& 
p 



COMMERCIAL TYPES 209 

blades, and balance pistons. The collars on the balance 
pistons which form the labyrinth packing are plainly visible. 
The increasing length of the blades of the third (exhaust) 
section is also very apparent. 

Besides the Westinghouse Machine Company of Pittsburg, 
Pa., other important manufacturers of Parsons turbines are the 
following : 

Allis-Chalmers Company, Milwaukee, Wis. 

C. A. Parsons & Co., Newcastle, England. 

Willans-Robinson Company, Rugby, England. 

Brown-Bo veri & Co., Baden, Switzerland, and Mannheim, 

Germany.* 
British Westinghouse Company, Manchester, England. 

The Allis-Chalmers steam turbine is a reaction type which 
differs from the original Parsons machines principally in manu- 
facturing details intended to remove some of the operating diffi- 
culties of the older designs. An innovation in the design of these 
turbines is in the arrangement and construction of the balance 
pistons. In the older types of reaction turbines the three balance 
pistons were put at the high-pressure end of the turbine. Some- 
times, however, there was difficulty with this construction, as 
the largest or low-pressure piston in large turbines was of com- 
paratively large diameter, so that an inner web was required in 
its construction. This web sometimes tended to warp so as to 
bring the "dummy" or baffle rings of the labyrinth construction 
on these pistons into contact with those attached to the casing. 
To overcome this difficulty the largest balance piston has been 
placed at the low-pressure end of the rotor behind the last row 
of blades. In this location its effective area starts from a smaller 
inner diameter, so that the required area can be obtained with a 
smaller outer diameter. 

Fig. 102 represents diagrammatically an Allis-Chalmers 

* A 24,000-horsepower steam turbine has been constructed at the Mannheim 
works of Brown-Boveri & Co. for the Krupp steel works and blast furnace plant at 
Rheinhausen. It is probably the largest turbine yet ordered for stationary service. 

The governing and overload valve designs of Brown-Boveri & Co.'s turbines 
are described and discussed on pages 292 and 305. 



210 



THE STEAM TURBINE 




3^3 




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(211) 



212 



THE STEAM TURBINE 



design of Parsons turbine: There are three sections of the 
rotor — H, J, and K — and three corresponding balance pistons, 
L, M, and Z. The construction of the rotor of one of these 
turbines is shown in Fig. 209. Steam admission valves are 
shown as in the usual Parsons designs. The valve D admits 
steam to the high-pressure end of the turbine and is always under 
the direct control of the governor. The second valve, V, called 
the overload valve, is opened only when the turbine must be 
operated at overload or non-condensing when the condenser 
equipment is out of service (see page 307). At C the main steam 
pipe enters the steam-chest and the exhaust is at G. Main bear- 
ings are at A and B. 

A 5500-kilowatt Allis-Chalmers turbine-generator is illustrated 
in Fig. 103. 

Governors and Low-Pressure Turbines. The various methods 
for governing Parsons turbines and the designs of low-pressure 
steam turbines of the Parsons type are discussed in Chapters 
VIII and IX. 











































































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3000 4000 

Rated Full Load Output— Kw. 



Fig. 104. Approximate Steam Consumption of Any Size of Parsons Turbine. 



Economy Curves. Fig. 104 shows fair average values of the 
steam consumption of good designs of Parsons turbines for 165 
pounds per square inch absolute steam pressure, 28 inches 
vacuum, and o° F. superheat. American Parsons turbines, until 
recently, were not made in smaller sizes than 400 kilowatts. 
Typical tests and load curves of 300, 500, and 1000 kilowatt 
Parsons turbines are given on pages 360 and 361. 

The curves in Fig. 105 are based upon the results of tests of a 
Westinghouse-Parsons steam turbine of standard construction. 



COMMERCIAL TYPES 



213 



It is stated by the manufacturers that the performance as shown 
by these curves is typical of machines of this type. 

The diagonal lines or "Willans lines" in the figure show the 
total water weighed or steam condensed per hour at various loads. 
The curves or "water rate curves" show the variation in water, 
or more correctly, in steam consumption per horsepower-hour at 



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1000 1600 

LOAD IN KILOWATTS 



Fig. 105. Typical Economy Curves of a 1000-Kilowatt Westinghouse-Parsons 

Steam Turbine. 



various loads, that is, the "water or steam rate" of the turbine. 
Each " water rate curve" corresponds to a "Willans line" — the 
upper curve to the upper line, the lower curve to the lower line, etc. 
Operating conditions of these tests are : 

(1) Condensing — saturated and superheated steam (ioo°F.) 

(2) Non-condensing — saturated and superheated steam (ioo° 



F. 



(3) One-quarter rated load to 100 per cent, overload. 



214 



THE STEAM TURBINE 



In the two overload tests the operation of the automatic 
secondary or overload valve may be observed. As before noted, 
it comes into action at a definite predetermined load as indicated 
by a bend in the water line. With the aid of this valve the best 
economy of the turbine is secured throughout the range of normal 
loading, while large overload capacity is available when desired, 
although at slightly decreased efficiency. When the secondary 
valve, however, has come fairly into action, the efficiency under- 
goes gradual improvement, as shown by the reversal of curvature 
of the curves of steam consumption. 



16 

15.5 

15 

14.5 

14 

13.5 



2 13 



12.5 



12 



































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Effect of Vacuum and Superheat 
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1500 K.W. Turbine Full Load 


















































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26 27 

Vacuum - Inches 



20 40 60 80 100 120 140 
Superheat Deg. F. 

Fig. io6. Curves of Steam Consumption of a 1 500-Kilowatt Westinghouse 
Turbine with Varying Vacuum and Superheat. 

A turbine designed for condensing work will not operate non- 
condensing with quite as good economy as if designed to exhaust 
against atmospheric pressure. That this economy is, however, 
excellent is shown by the upper pair of curves. The water rate 
is somewhat less than double the condensing water rate. 

Fig. 106 illustrates graphically the effect of vacuum and super- 
heat on the steam consumption of a 1500-kilowatt Westinghouse 
turbine. The percentage change in the steam consumption is 
said to be about the same for all sizes. 



COMMERCIAL TYPES 215 

THE WESTINGHOUSE "IMPULSE AND REACTION" DOUBLE- 
FLOW TURBINES. 

The double-flow principle has been adopted recently for the 
design of large sizes of Westinghouse turbines largely for mechan- 
ical reasons — primarily to avoid the end thrust which is an impor- 
tant factor in all reaction types. In small machines, however, 
the double-flow principle does not have the same advantages as 
in the large machines. It is very obvious that the economy of 
two small machines is not nearly as good as one of twice the 
capacity. With large machines, however, the change in econ- 
omy is not nearly so great when the capacity is doubled. This 
fact is well illustrated by the curve in Fig. 104. Test results as 
shown for large sizes are of the combined " impulse and reaction' ' 
type. 

Fig. 107 illustrates a Westinghouse double-flow turbine with 
an impulse element. In its essential parts this turbine consists 
of a set of nozzles, an impulse wheel with two velocity stages, one 
intermediate section, and two low-pressure sections of Parsons 
blading. Steam enters the turbine through an opening in the 
lower half of the casing,* from which it is piped directly to the 
nozzle block shown at the top of the figure. Steam escapes from 
these nozzles f at a high velocity to impinge on the impulse 
blades. The casing around the impu se wheel is made of suffi- 
cient size to permit a good distribution of the steam, so that it 
will enter the intermediate Parsons sect on evenly around the 
entire circumference of the rotor. After the steam has passed 
through the intermediate section it divides along two separate 
paths. One-half enters the left-hand section of the low-pressure 
Parsons blading and the other half passes through the interior of 

* Advantage of steam entering the lower half of the casing is that the top can 
be removed without disturbing the piping supplying the turbine. 

t These nozzles are made non-expanding. It has been shown that non-expand- 
ing nozzles give higher efficiencies than expanding nozzles with steam at less than 
about 70 pounds gauge pressure. (See footnote, page 55.) The designers of these 
turbines have recognized that there are nozzle losses due to under-expansion in a 
diverging or expanding nozzle when the steam is throttled at light loads. 



2l6 



THE STEAM TURBINE 



the rotor shell which forms the connecting passage to the right- 
hand low-pressure section. Arrows indicate in the figure the 
passage of the steam through the shell. When the steam is dis- 
charged from the last rows of low-pressure blades, it passes into 




the exhaust pipes — of which there is one at each end — and 
then to the condenser. 

As there is practically no expansion in the impulse blades, these 
blade areas are made to increase only in proportion to the reduc- 
tion in steam velocity in each row of moving blades. 



COMMERCIAL TYPES 217 

As the same pressure exists on both sides of the impulse wheel 
disk, this is not subjected to any end thrust, and requires no 
balancing. The small thrust due to the difference of pressure 
between the inlet and outlet of the Parsons intermediate section 
is accurately equilibrated by a "dummy" or balance piston, of 
moderate dimensions, located between the impulse wheel and 
the right-hand low-pressure section. The thrusts in the low- 
pressure sections are in opposite directions, and are therefore 
balanced. With these arrangements it is possible for the entire 
turbine to run in perfect equilibrium under all conditions of 
vacuum, pressure, and load. It is, of course, necessary to provide 
means for accurately fixing the axial position of the rotor, and 
for this purpose an adjustment bearing, shown at the right-hand 
end of the shaft in Fig. 107, is provided. It consists of a number 
of collars turned in the steel shaft, into which fit corresponding 
brass rings fixed in the adjustment blocks. The upper and 
lower halves of the adjustment bearing may be moved by means 
of micrometer screws, thus permitting the axial position of the 
rotor to be accurately known at all times. 

All double-flow cylinders are made in two parts, the upper and 
lower halves each being a one-piece casting. The design is 
symmetrical throughout, devoid of longitudinal flanges except 
those at the center required for bolting the two parts together. 
The castings are first rough-bored, after the flanges have been 
planed and drilled, and are then "seasoned" with high-pressure 
steam for a number of hours to remove any local casting stresses 
in the metal. They are then given the finishing cut and assembled 
with the boring bar running in the bearing housing so as to 
insure a concentric bore. Manholes are provided at each end 
of the cylinder to permit access for interior examination, and 
auxiliary relief valves are fitted in each of the manhole covers to 
prevent the pressure in the exhaust passages from rising to a 
dangerous point in case of failure of the condensing apparatus 
or the sticking of the atmospheric relief valve in the exhaust 
piping as otherwise dangerous pressure would result in the casing. 

A Y-connection, fitted with two corrugated copper expansion 



2l8 



THE STEAM TURBINE 



joints located below the turbine, connects the separate exhausts 
to the main exhaust pipe. These expansion joints provide for 
the expansion and contraction of the turbine casing. 

The rotating element of the turbine is built up of five cast-steel 
parts, in addition to the shaft. As may be seen in Fig. 107, 
these are the three Parsons blading supports, the impulse section 
and a dished plate. It is stated that the shaft carries its load 




Fig. 108. 7500-Kilowatt Westinghouse Turbine. 



(the weight of the rotor on the end supports) at one-third the 
distance from the points of support, so that this design allows a 
lighter shaft than would be required for distributed loading, and 
the consideration of deflection is practically eliminated. This 
built-up part of the rotor is rigidly attached to the shaft only at 
the right-hand support, and the opposite end is fitted with a 
bronze bushing surrounding the shaft, so as to permit the rotor 



COMMERCIAL TYPES 219 

to move axially, without appreciable resistance, under any dif- 
ferential expansion of shaft and rotor.* The impulse section 
consists of a flanged cast-steel disk forced on the body carrying 
the intermediate Parsons blading. The flange of this disk is 
grooved at the base and forms the dummy piston for balancing 
the thrust of the intermediate Parsons section. Fig. 108 is a 
half-tone illustration of a 7500-kilowatt Westinghouse double- 
flow turbine. 

The rotor of a 6000-kilowatt double-flow turbine is shown in 
Fig. 109. Details of the arrangement of nozzles and blades are 




Fig. 109. Rotor of a Westinghouse Double-Flow Turbine. 

shown in Fig. no. It is seen that the nozzle block is a casting 
quite separate from the turbine casing. As it receives steam 
from the governor valve, high temperature steam is restricted 
to a comparatively small casting which is made free to expand 
and contract with changes of temperature. 

A new type of shaft coupling for Westinghouse turbines is 
illustrated in Fig. in. 

Westinghouse Emergency Speed Limit. A very interesting 
mechanism is provided with Westinghouse turbines for shutting 
off the steam supply in case the governor fails to act and a 
dangerous speed might be attained. Details of this mechanism 
are shown in Figs. 112a and 112b. In its essential elements it 

* This is an ingenious design but a more recent construction of Westinghouse 
rotors is shown in Fig. ii2g, where the end sections of the rotor are cast integral 
with the sections of the shaft, as required at each end. 



220 



THE STEAM TURBINE 




mfftf j 



Fig. iio. Westinghouse Nozzle Block, Showing Arrangement of Nozzles and 

Blades. 



to ^■■I'pN ^'" 






Fig. hi. Westinghouse Shaft Coupling. 






COMMERCIAL TYPES 



221 



consists of a " weight pin " P, placed diametrically at right angles 
to the axis of the shaft, in a cylindrical "body" screwed on the 
main turbine shaft at the high-pressure end. Centrifugal force 
tends to drive this pin away from the center and through the 
loosely fitting collar N. This force is resisted, however, by 
the " weight spring " shown around the pin in the figures. 
The strength of this spring can be adjusted by means of the 




COVER 

TRIGGER CAM BUSHING 

TRIGGER 
TRIGGER CAM 

TRN 
WEIGHT SPRING RETAINER LOCH 



BALANCING BLOCK Pi 
BALANCING BLOCK 
BODY H 




WEIGHT SPRING RETAIN 
BODY LOCKING SCR 



WEIGHT SPRING RETAINER LOCK SPRIi 



—VALVE LEVER 
VALVE LEVER PIN - 
I VALVE LEVER PLATE 

., TRIGGER PLATE 
—TRIGGER STOP 
TRIGGER 

— VALVE BODY COVER 

-VALVE BODY (UPPER) 

-VALVE SPRING ADJUSTING SCREW 

-VALVE SPRING RETAINER(UPPER) 

-VALVE SPRING 

-VALVE SPRING RETAINER(LOWER) 

—VALVE BODY (LOWER) 

r ALVE 



Fig. 112a. Phantom View of Westinghouse Emergency Speed Limit. 



collar N, which is provided with a screw thread. Such adjust- 
ment determines the speed at which the centrifugal force over- 
comes the spring and forces the pin outward to engage with a 
trigger cam L. This cam is rigidly attached to one end of a short 
shaft S, which carries at its other end a trigger H. A small plate 
at the bottom of the valve lever C is supported normally at one 
end on the trigger H and at the other end on a screw provided 
for adjusting the spring on the auxiliary steam valve E. 



222 



THE STEAM TURBINE 



If the speed of the turbine should become higher than the limit 
for which the "weight spring" is set, the pin P is forced out to 
engage with the cam L, which in turn moves the trigger H away 
from the valve lever plate which it supports. In this way the 
valve E is opened because the tension in the spring on its spindle 




Fig. 112b. Drawings of Westinghouse Emergency Speed Limit. 



is released. There is always high-pressure steam on the upper 
side of the valve E, and when it is removed from its seat this 
steam rushes through a pipe connecting the lower side of the 
valve to a small steam cylinder at the side of the main steam pipe. 
A short rod attached to a piston in this cylinder is moved by the 
steam pressure to strike a trigger which releases and closes the 
emergency valve on the main steam pipe. 



COMMERCIAL TYPES 223 

Advantages of the Westinghouse Double-flow Type. In large 
capacities the following advantages are claimed for the double- 
flow type over the usual Parsons designs : 

(1) Reduction in size and weight due to higher permissible 
speed. 

(2) Almost negligible end thrust. 

(3 ) Blades and casing are not exposed to steam at high temper- 
atures. 

(4) Large volume per pound of steam at the admission to the 
first Parsons section avoids the use of very short blades. 

(5 ) Only one balance piston is required and this is of relatively 
small diameter. 

(6) Exhaust connections are considerably reduced in size, due 
to divided flow. 

(7) The impulse element is well suited to high pressure and 
superheat, and by this modification the shaft length is 
reduced nearly 50 per cent. 

An exact reproduction of a section drawing of a Westinghouse 
double-flow low-pressure turbine rated at 1000 kilowatts is 
shown in Fig. 184 in Chapter IX. 

The following figures 112c, d, e, and f show designs used by 
the Westinghouse Machine Company for 25,000-kilowatt turbine- 
generators to operate at 200 pounds per square inch absolute 
steam pressure, 29 inches vacuum and 200 F. of superheat. 
Fig. 112c shows the double-flow turbine designed for these con- 
ditions and operating at 1500 r.p.m. Fig. ii2d shows a similar 
design operating at 750 r.p.m. A tandem compound arrange- 
ment operating at 750 r.p.m. is shown in Fig. ii2e. In this last 
design it will be observed that the high-pressure portion is of 
the ordinary single-flow arrangement while the low-pressure 
end is made double-flow. The combined unit is connected to 
a single 25,000-kilowatt generator. A cross-compound turbine 
arrangement, Fig. ii2f, with the high-pressure portion operating 
at 1500 r.p.m. and connected to a 12,500-kilowatt generator is 
shown here together with a low-pressure portion which is of the 
double-flow arrangement and operating at 750 r.p.m. The low- 



2 24 



THE STEAM TURBINE 



pressure portion is also connected to a 12,500-kilowatt generator. 
These figures show the comparison to scale of the arrangements 
described above. 

All of these designs would give very excellent economy, and the 
choice of the unit would depend primarily on the two factors of 
first cost and economy, assuming that in each case the relia- 
bility for continuous operations is the same. 





Fig. ii2c. Double-flow Reaction Turbine Designed for 1500 r.p.m. 

A close study of the four arrangements indicates that the 
double-flow turbine at 1500 r.p.m., direct connected to a single 
generator (Fig. ii2g), is the cheapest construction. The large 




Fig. i 1 2d. Double-flow Reaction Turbine Designed for 750 r.p.m. 

areas required in the low-pressure stages of this turbine make 
high velocity and long length of blades essential, with the 
necessity of careful designing to properly take care of the stresses 
due to centrifugal force in the low-pressure end. 



COMMERCIAL TYPES 



225 



The most economical combination of the four is the cross- 
compound reaction turbine with the high-pressure portion run- 
ning at 1560 revolutions and the low-pressure portion at 750 






Fig. 112c " Tandem-compound " Reaction Steam Turbine. 

revolutions. With this arrangement the highest efficiency is 
obtained, because the method of combining the unit into high- 
and low-pressure cylinders, running at 1500 and 750 r.p.m., 





Fig. ii2f. " Cross-compound " Reaction Steam Turbine. 

gives the condition for best blading proportions throughout the 
turbine without departing from established standards of prac- 



2 26 



THE STEAM TURBINE 



tice. The construction, however, is considerably heavier than 
the single unit of the double-flow type (Fig. ii2g) and is more 
costly to construct and install. At powers and speeds attain- 
able with single alternating-current units of, say, below 15,000 
kilowatts' capacity, the double-flow turbine will be nearly, if not 
equal, to the cross-compound " straight reaction" turbine under 
the same operating conditions, and any difference in efficiency 
would probably be offset by the lower first cost of the double-flow 
machine. Taking into consideration both first cost and effi- 
ciency, between 10,000 and 40,000 kilowatts' capacity, the double- 
flow machine is undoubtedly the proper type of construction. 




Fig. ii2g. Section of Combined Impulse and Double-flow Reaction Turbine. 



The demand for turbine-generators is greatest between 4000 
to 15,000 kilowatts maximum rated capacity, within which range 
the double-flow machines of the combined impulse and reaction 
types of blading satisfactorily meet commercial conditions both 
with respect to cost and efficiency. There is, however, practi- 
cally the same to be said regarding turbines of the Curtis type 
having six to twelve pressure stages. 

Below 4000 kilowatts' capacity the turbines consisting of an 
impulse wheel followed by single-flow reaction blading, and a 
" straight " single-flow reaction turbine (Fig. 102, page 210, 
and Fig. H2i), represent the machines best suited for average 
operating conditions. In most cases the former would be pre- 



COMMERCIAL TYPES 227 

f erred; but when the speed is to be made particularly low the 
preference goes to the latter. 

Thus with a speed of 3600 r.p.m. driving a 60-cycle generator 
at, say, 500 kilowatts, the best design would be a combined 
impulse and reaction machine for best efficiency and lowest 
cost. If, however, the generator to be driven was for 25-cycle 
sendee, with an allowable maximum speed of 1500 r.p.m. and 
the same capacity the single-flow reaction turbine would be 
selected, providing reaction blading was to be used at all. 

In general, however, the application of the combined impulse 
and reaction turbine, consisting of an impulse element for the 
high-pressure portion and reaction blading for the low-pressure 
portion, is well adapted for complete expansion turbines over 
wide ranges of power and speed; and since the introduction of 




Fig. 112E. Relative Lengths of Rotors in Two Common Types. 

this type, a large proportion of the firms building steam turbines 
have utilized this construction, either with Parsons or Rateau 
blading following the Curtis impulse element in the high- 
pressure end. The principal advantage of this type of con- 
struction is the shortening of the machine without very much 
loss in efficiency, the elimination of balancing pistons with the 
avoidance of the very considerable leakage of steam through 
them, and the securing of high economies at light loads by the 
application of the method of governing by " cutting out nozzles" 
(see page 277), now applied to Westinghouse turbines of the 
combined impulse and reaction type. 



228 



THE STEAM TURBINE 






Fig. ii2h shows the relative lengths of the rotors of the latest 
design of Westinghouse turbine with combined impulse blading 
and single-flow reaction blading compared with the conventional 
single-flow Parsons type with " balance pistons " (see page 198). 
Nearly 50 per cent, in length is saved, making the difficulties 
due to springing and expansion of the casing and rotor relatively 
small. 

Another important consideration in choosing between the 
" straight " single-flow reaction and the combined impulse and 
reaction types is that the former is generally preferred for 




Fig. 112L Single-flow Reaction Turbine with Kingsbury's Thrust Bearing. 

moderate superheats and pressures, while the latter is selected 
when the superheats and pressures are high. 

Another important improvement in the construction of reac- 
tion turbines is shown in Fig. ii2i where the Kingsbury type of 
thrust bearing is applied. By this means the usual " balance 
pistons " required for the single-flow type are eliminated. 

A recent design of a 20,000-kilowatt " tandem " type of reac- 
tion turbine is shown in Fig. 112J. At the right-hand side is the 



COMMERCIAL TYPES 



229 




Fig. 112J. Tandem Reaction Turbine. 

high-pressure turbine and at the left the low-pressure. This 
figure shows more clearly the type shown also in Fig. 112c 
There are very few applications as yet for " land " service of 
this arrangement, although it is common for marine service. 




Fig. 112k. Relative Sizes of Steam Turbine-generators, from One Kilowatt to 
35,000 Kilowatts. 



Fig. 112k illustrates the relative sizes of steam turbine-gener- 
ators for capacities from one kilowatt for the smallest to 35,000 
kilowatts for the largest. The turbine is shown on the right- 
hand side and the generator on the left. 



230 THE STEAM TURBINE 

THE CURTIS TURBINE 

The Curtis steam turbine, of which the original patents were 
issued to C. G. Curtis about 1895, is manufactured by the 
General Electric Company at Schenectady, N.Y., and Lynn, 
Mass., the British Thomson-Houston Company at Rugby, 
England, and the Allgemeine Elektrizitats Gesellschaft at Berlin, 
Germany. 

As in the De Laval turbine, the steam is expanded in nozzles 
before reaching the moving blades, but the complete expansion 
from the boiler to the exhaust pressure occurs in this type usually 
in a series of stages or steps, as the steam passes through a suc- 
cession of chambers, separated from each other by diaphragms. 
The diaphragms and blade wheels of a four-stage Curtis turbine 
are shown by a section drawing in Fig. 113. Each chamber or 
stage contains usually one disk or blade * wheel. Steam at the 
admission pressure enters the first set of nozzles through the port 
A, where it expands to the pressure in the first stage and delivers 
a portion of its energy to the blades in the wheel F. The steam 
then expands again through a second set of nozzles in the dia- 
phragm C leading to a still lower pressure in the second stage, 
where it gives up a portion of the energy remaining to a second 
set of blades, and so on. In the very small units but one pressure 
stage is usually employed, but in the larger sizes from two to five 
are used. The general arrangement of the nozzles and blades 
in a single-stage Curtis turbine was shown diagrammatically in 
Fig. 39. It is typical of these turbines that there are always 
three or more rows of blades following each set of nozzles, and 
at least one row is stationary. These stationary blades are 
technically called intermediates. There is practically no expan- 
sion in the stationary blades; the object of the several rows of 
blades is only to reduce the velocity, and for a given blade speed 

* The terms vane, blade, and bucket are often used interchangeably. 
Common practice, however, seems to apply blade to the Parsons turbine, and 
bucket to the Curtis, De Laval, and those of the Pelton type. In order, however, 
that the notation may not be confused, the term blade will be used in connection 
with Curtis as well as other types. 



COMMERCIAL TYPES 



231 







Fig. 11 x. Section of a Four-stage Curtis Turbine Showing Diaphragms and 



Blade Wheels. 



232 



THE STEAM TURBINE 



the steam velocity is reduced per pressure stage in proportion to 
the number of rows of moving blades. Each pressure stage is 
said then to have as many velocity stages as there are rows of 
moving blades. It must be noted, however, that, unlike the case 
of the Parsons turbine, the steam expands only in the nozzles, 
and the pressure is practically the same on both sides of any row 
of blades. 

Nozzles. The nozzles are generally rectangular in cross-sec- 
tion, with " rounded corners." They are grouped closely to- 
gether, being cast either integral with the diaphragms or in 
separate plates (Fig. 1 14) , which in assembling are bolted to the 
diaphragms. The number of nozzles is proportioned to the 




Fig. 114. Nozzle Plates of Curtis Turbines. 



power required and the degree of expansion used; in some cases, 
at least in the low-pressure stages, they extend completely around 
the diaphragm, making a continuous band of steam around the 
circumference. This method has the advantage of reducing 
blade rotation losses to a minimum, as explained in Chapter V. 
Steel, bronze alloys, and cast iron are employed for making the 
nozzles of Curtis turbines. 

Wheel Disks and Blades. The blade wheels are usually made 
of forged steel disks similar to Fig. 216, which increase in thick- 



COMMERCIAL TYPES 



233 



ness as they approach the hub, but in larger sizes the construc- 
tion shown in Fig. 113 is often employed. In some very small 
turbines the blades are cut in the solid rim by special machines, 
while others have drawn or rolled blades which are cast into 
segments (Fig. 115) of bronze alloy designed to be riveted to the 
rim. A dovetailing meth- 
od similar to Fig. 63 is 
now generally preferred 
to the method of inserting 
the blades by casting. 
The fixed blades, or inter- 
mediates, are also either 
cut or cast in segments 
(Fig. 116), and are fas- 
tened by bolts to the interior of the casing as shown in Fig. 57. 
These intermediates cover only the portion of the circumference 
upon which the belt of steam delivered by the nozzles can im- 
pinge. To make the blades more rigid, thin bands or shroud 
rings are riveted in segments to projections on their ends. 




Fig. 115. Curtis Moving Blade Segments. 




Fig. 116. Curtis Intermediate Blade Segments. 



The wheels of a four-stage Curtis turbine are shown in Fig. 
117. There are two rows of blades on each wheel, so that in this 
design there are two velocity stages in each pressure stage. The 
shroud rings on each row of blades are plainly visible. 

Shafts and Bearings. The smaller sizes of Curtis turbines 
have horizontal shafts with standard bearings, as devices for 
flexibility are unnecessary at the speeds employed. The larger 



234 



THE STEAM TURBINE 




COMMERCIAL TYPES 



235 



sizes, however, are sometimes built with a vertical shaft supported 
on a step bearing, shown at the bottom of Fig. 118, which is 
supplied with oil or water under pressure,* the shaft thus revolv- 
ing on a film of liquid. The small disk D is attached by dowels E 
to the main shaft. The bearing is between the stationary plate 
C and the disk D. This vertical shaft arrangement was formerly 
one of the special characteristics of the large sizes of Curtis tur- 




y//////////, 
steer?? 

Supply 




0/7 Drain 
fZ—O/7 dupp/y 

Fig. 118. Step Bearing for a Vertical Curtis Turbine. 



bines, and produces a very compact design. The direct-con- 
nected electric generator is mounted immediately above the 
turbine, as shown in Fig. 119, which is a section of a 9000- 
kilowatt Curtis turbine-generator. 

Fig. 120 is a " phantom " view of a 300-kilowatt Curtis turbine- 
generator, showing the wheels, armature, and couplings as if the 
turbine casing and generator frame were transparent. 

Curtis units are manufactured from 15 kilowatts (about 20 

* Water pressure is usually 500 to 600 pounds per square inch. 



236 



THE STEAM TURBINE 




Fig. 119. Section of 9000-Kilowatt Vertical Curtis Turbine-Generator. 



COMMERCIAL TYPES 



237 




Fig. 120. Phantom View of a Curtis Turbine Showing Wheels, Armature, and 

Couplings. 




Fig. 121. Ring Type of Emergency Stop. 



2 3 8 



THE STEAM TURBINE 



horsepower) at 3600 to 4000 revolutions per minute to as high 
as 9000 kilowatts (nearly 12,000 horsepower) at about 750 
revolutions per minute, the general application being to direct- 
connected electric generators for power or lighting purposes. 

Emergency Valve. Since a steam turbine can accelerate at 
a rapid rate and this increase in speed is not easily perceptible. 




Fig. 122. Emergency Stop Valve. 



it is important that all these machines be equipped with simple 
speed limiting devices which are operated automatically in emer- 
gencies. The device shown in Fig. 121 consists of a steel ring 
(13) placed around the shaft between the turbine and the gener- 
ator. This ring, which is held in place by stud bolts (4), is 
placed in a slightly eccentric position, and the centrifugal force 
due to this unbalancing is counteracted by a helical spring (n). 



COMMERCIAL TYPES 



2 39 



When the speed increases, the centrifugal effort overcomes the 
spring and the ring moves into a still more eccentric position as 
indicated by the dotted lines. In this position the ring strikes a 
bell-crank lever, which trips, by means of a simple auxiliary 
mechanism and the tension rod L (Fig. 122), the throttle valve 
on the main steam supply pipe. The rod L is connected to the 
crank D, which operates to release the spring S, pulling up the 
gear and throwing out the hook G, which holds the valve open. 
When released by this emergency ring mechanism, the valve 
descends upon its seat with a very positive motion due to its 




Fig. 123. Details of "Spring Type" of Emergency Stop. 

own weight and the unbalanced pressure on the area of the 
valve stem. 

Fig. 123 shows a little different arrangement for tripping the 
valve. The free end of a spiral spring is thrown out by centrifu- 
gal force and strikes a bell-crank lever in very much the same 
way as the ring does. The emergency valve is opened by means 
of the hand wheel shown at the bottom of the figure. 

No turbine should be kept in operation unless it is known that 
this speed limiting device is in reliable condition. 

Governor. Curtis turbines are governed by a method com- 
monly known as "cutting out nozzles." By this method the 



240 



THE STEAM TURBINE 



number of nozzles which are open for the discharge of steam is 
regulated according to the requirements of the load. This method 
is described and typical Curtis governors and valve gears are 
illustrated in Chapter VIII. 




FlG. 124. 25-Kilowatf Curtis Turbine-Generator. 

Small Turbines. Fig. 124 shows a 25-kilowatt Curtis turbine 
and generator suitable for lighting a factory. The whole set 
occupies very little space compared with that required for a 
reciprocating engine. The shaft, armature, and turbine wheel of 
this set are shown separately in. Fig. 125. One of the latest and 




Fig. 125. Wheel, Shaft, Armature, and Commutator of a Small Curtis Turbine, 

most efficient designs of blade or bucket wheels for Curtis tur- 
bines with two pressure stages is shown in Fig. 125a. The 
disks or wheels in the two stages are of the same diameter but 
the much greater blade length toward the low-pressure end 



COMMERCIAL TYPES 



2 4 l 



makes the actual over-all diameter at that end considerably 
larger than at the high-pressure end. 




FlG. 125a. Latest Construction of Blade Wheels of Curtis Two-stage Turbines. 

The most recent improvement in valve gears on Curtis tur- 
bines is shown on the turbine illustrated in Fig. 125b. The 




Fig. 125b. Horizontal Curtis Steam Turbine with Latest Steam-operated Valve 

Gear. 



242 



THE STEAM TURBINE 



centrifugal governor is placed at the upper end of a vertical shaft 
between the turbine and generator, which is driven by worm 
gearing from the main shaft of the turbine. The motion of the 
main governor is transmitted to the valve gear * by means of 
levers and rods, which operate a small pilot valve, controlling 
the admission of steam to a steam cylinder at the upper end of 
the valve mechanism. That is, the pilot valve serves to admit 
steam either above or below the piston in the steam cylinder. 
The piston rod extends into the steam chest and on this rod 
are mounted a series of spiders, which engage a corresponding 
series of annular double-seated admission valves. The spiders 
on the valve rod are arranged so that the valves are lifted from 
their seats in sequence as the rod is raised by the steam cylinder 
under control of the pilot valve. As each of the valves is lifted 
from its seat, steam is admitted from the central space within 
the annular valves to passages leading to the turbine nozzles. 

Correction Curves. Typical curves showing the variation in 
steam consumption of a 500-kilowatt Curtis turbine, due to in- 
creasing superheat and vacuum, are shown in Figs. 126 and 127. 



:*8 






i! : 











































































































X 


s 
















































































S 


s 


























■s 



























JW 10 60 80 .100 120 110 160 180 200 220 
Superheat Degs.Eahr. 

Fig. 126. 



.sis 






24 

Vacuum 



Fig. 127. 

Curves Showing the Effect of Superheat and Vacuum on the Steam Consumption 
of a 500-Kilowatt Curtis Turbine. 



Such curves become most useful, however, when they are re- 
duced to equivalent percentages like those for De Laval turbines 

* This valve gear is described more completely with the help of illustrative 
figures on pages 289 and 291. 



COMMERCIAL TYPES 



243 



shown on pages 192 and 193. In Chapter VI the correct method 
for making this transposition was explained. 

Steam Consumption. Fig. 128 is a curve to show approxi- 



If 



3000 4000 

Rated_Full Load Output — Kw. 



5000 



Fig. 128. Approximate Steam Consumption of Any Size of Curtis Turbine 
with 165 Pounds per Square Inch Absolute Pressure, 28 Inches Vacuum, and 
no Superheat. 

mately the steam consumption of any size of Curtis turbine at the 
rated full load. All the data for this curve were corrected by 
using percentage curves like those referred to above, which served 
to reduce the conditions of the various tests to assumed conditions 
of 165 pounds per square inch absolute steam pressure, 28 inches 
vacuum, and no superheat. To get sufficient data for this 
curve it was necessary to include some tests made with com- 
mercial loads, making its values probably a little higher than 
they would be if all the tests had been run with a constant load. 
Analysis of Losses in a Curtis Turbine. Steinmetz has calcu- 
lated the energy distribution in a typical two-stage Curtis tur- 
bine and has given the results in the diagram in Fig. 129. 

WESTINGHOUSE IMPULSE TURBINES. 

Still another type of steam turbine intended particularly for 
small capacities has been developed by the Westinghouse Ma- 
chine Company, as illustrated in Figs. 130 and 131.* Machines 
of this type are suitable for a capacity as low as one kilowatt. 
By this construction it is possible to secure with the use of only 

* Turbines of this type are known abroad as " Electra " designs. In these 
foreign turbines the direction of flow of steam is different in that it is radial. 



244 



THE STEAM TURBINE 









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Fig. 129. Analysis of the Losses in a Turbine with Three Velocity Stages in 
Each of Two Pressure Stages. 




Fig. 130. Westinghouse Impulse Turbine (with one reversal). 



COMMERCIAL TYPES 



245 



one row of moving blades an effect similar to the velocity stages 
in a Curtis turbine with only one row of moving blades or 
buckets as illustrated in Fig. 130. This design is suitable for 
the pressure drop in non-condensing operation. The arrange- 




F™£* 



Fig. 131. Westinghouse Impulse Turbine (with two reversals). 

ment shown in Fig. 131 has two reversals of the steam and is 
suitable for condensing operation. 




Fig. 131a. Three Sizes of Small Westinghouse Impulse Turbines. 

The advantages of this construction are that it is essentially 
simpler than the De Laval in the ehmination of speed-reduction 
gears, and requires a very much smaller number of blades 
than the Curtis type. 



246 



THE STEAM TURBINE 




COMMERCIAL TYPES 



247 



RATEAU TURBINES. 



Professor Rateau of Paris is also a pioneer in the development 
of a well-known type of steam turbine. His first experiments 
were made with a turbine having a single impulse wheel; but he 
soon abandoned this type in favor of a multiple wheel construc- 




Fig. 132. Diagrammatic Representation of Four Stages of a Rateau Turbine. 



tion. The Rateau turbine is often called " multicellular," 
meaning that it consists of a large number of " cells " or 
pressure stages of which the separating walls are diaphragms 
similar to those in a Curtis turbine. The principle of the 
Rateau turbine is illustrated by the section drawing in Fig. 132, 
which shows diagrammatically four stages. Essentially the 



248 THE STEAM TURBINE 

Rateau type differs from that of Curtis in that it has a much 
larger number of pressure stages or "cells" but no velocity 
stages. There is therefore only one row of blades in each stage. 
Except for the fact that turbines with simple disk wheels can be 
operated at higher blade speeds than reaction turbines of the 
drum type (Parsons), making the efficient utilization of steam at 
higher velocities possible, the Rateau and the Parsons types 
would require the same number of stages. Rateau turbines 
have from 20 to 40 stages respectively, depending on whether 
they are for non-condensing or for condensing service. For 
given blade-speed, steam pressure, and superheat, the number of 
stages increases, although not proportionally, as the exhaust 
pressure is reduced. 

Nozzles and Diaphragms. Annular nozzles are set in each of 
the diaphragms between the stages. Because of the large num- 
ber of stages, the pressure drops are very small, so that the 
nozzles are made with a uniform cross-section along their length; 
that is, they are non-expanding. To allow for the increased 
volume of the steam as it expands, in almost all the other types 
of impulse turbines the nozzles are made with at least somewhat 
larger radial width for the lower pressures. In Rateau turbines, 
however, the same increased nozzle area is secured by increas- 
ing only the arc or part of the circumference occupied by the 
nozzles. In the last stages, then, where the entire circumference 
of the diaphragm is made use of, a complete annular jet 
results. 

Rateau nozzles are arranged in groups very much like the 
Curtis nozzle plate shown in Fig. 114. Diaphragms of several 
sizes of these turbines are shown in Fig. 133. Several groups of 
nozzles can be seen in each diaphragm. At the high-pressure 
end of the turbine there are only a few groups (usually about 
three), but in each succeeding stage there is a greater number. 
Because the steam discharged from the blades is carried along a 
short distance by the rotation of the wheel, a portion of each 
group of nozzles is located a little in advance of the preceding 
set. 



COMMERCIAL TYPES 



249 



One of the advantages claimed by Rateau for multi-stage 
types over those in which the steam is admitted around the whole 
periphery in all the stages, is that since the volume of the steam 



y^'-"- 


1 


: 





Fig. 133. Diaphragms of a Rateau Turbine Showing Nozzles and a 
Shaft Packing with "Water Grooves." 



at the admission end is small, the blades, in the Parsons type for 
example, have necessarily a small radial height, so that there is 
more friction due to the passage of steam than where the steam 
spaces are larger and the volume of the steam is large in pro- 
portion to the surface of the blades. 



250 



THE STEAM TURBINE 




COMMERCIAL TYPES 



25 1 



Description. Fig. 134 is a section of a 500-horsepower Ridg- 
way-Rateau turbine of seven stages. It will be observed that 
this turbine is a remarkably simple design. This turbine is de- 
signed to operate at 2400 revolutions per minute. In this figure 
the main steam pipe is shown at the bottom of the casing. 
From this pipe the live admission steam is discharged into the 
cored passages supplying the first stage nozzles after passing 
through the throttle valve controlled by the governor. The 
exhaust pipe leading to the condenser is cast integral with the 




Fig. 135. Cast-in Nozzles of Ridgway-Rateau Turbines. 



base at the right-hand end of the casing. Carbon packings in 
bushings of anti-friction metal are fitted in the diaphragms 
where the shaft passes through them. Steam leakage through 
these packings is reduced to a very small amount. 

A Moore steam turbine which differs from the conventional 
Rateau type in having two velocity stages in the first pressure 
stage is shown in Fig. 137. 

Wheel Disks. Typical Rateau disks are shown in Fig. 136. 
Details of construction are shown better, however, in Fig. 132. 
A shroud ring is fitted around the blades as illustrated in the 
latter figure. The blades resemble those used in De Laval and 
Curtis turbines except that they have a flat projection at the root 
which is provided to fasten them to the flange of the disk by 
riveting. The holes shown in the disk in Fig. 132 were drilled 



252 



THE STEAM TURBINE 



for balancing. Fig. 136 shows a group of Rateau disks assembled 
on the turbine shaft. 

Manufacturers. Rateau turbines are constructed by the 
pioneers, Sautter, Harle & Co., at Paris, by the Maschinenfabrik 
Oerlikon in Switzerland and by many other companies in Europe. 
American types are made by the Ridgway Dynamo & Engine 
Company of Ridgway, Pa., and the Southwark Foundry & 
Machine Company of Philadelphia. Rateau designs are fre- 
quently used in combination with other types, as for example 




Fig. 136. Ridgway Disks Assembled on the Turbine Shaft. 

when Curtis blading is used for the first stage and Rateau blad- 
ing for the remaining stages (see Fig. 137 and the example, 
pages 102-111). 

Low-Pressure Rateau Turbines are extensively used in Europe 
to operate with the exhaust steam from rolling mill and mine 
engines. Professor Rateau has designed a steam accumulator 
(Fig. 184) for application in such cases where the steam supply 
is intermittent. It is described in Chapter IX in the discussion 
of low-pressure steam turbines. 



COMMERCIAL TYPES 



253 






^3 




m 



H H 



p- 254 



COMMERCIAL TYPES 



255 



WILKINSON TURBINES. 

Rateau turbines are governed by throttling the steam pressure 
by means of valves controlled by the governor. Mr. James 
Wilkinson has invented a system of governing steam turbines 
(see page 302) which is intended to be equivalent to the Corliss 
" cut-off " governing of reciprocating engines. He has applied 
this method of governing, together with some other unique 
features, to steam turbines of the Rateau type which were made 
at the Corliss Engine Works, Providence, R. I. A Wilkinson 
turbine-generator rated at 100 kilowatts for non-condensing 
service (six stages) is shown by side and end sections in Figs. 140 
and 141. It will be observed that in this design the diaphragms 
are " dished " as in Curtis turbines, while the disks are flat. The 
disks are made of forged steel, but the blades are bronze castings 
which are filed to a sharp edge on the side where the steam enters. 

Stage Packing. To prevent the leakage of steam between the 
diaphragms and the shaft (stage leakage) , which in some impulse 




Fig. 142. Wilkinson Labyrinth Stage Packing. 



turbines is a considerable loss — often 10 to 20 per cent. — a 
very ingenious system of steam packing has been devised. A 
drawing illustrating this system is shown in Fig. 142. By this 
device, steam containing a large amount of condensation is dis- 
charged into grooved packings between the diaphragms and the 



256 THE STEAM TURBINE 

shaft through ducts drilled into the hubs of the disks. This wet 
steam is taken from a part of the labyrinth packing at the high- 
pressure end of the turbine — through which there is always 
some leakage of steam — and is conducted in the ducts shown in 
the figure, which are arranged so that the steam discharged into 
a diaphragm packing is at a slightly higher pressure than that on 
either side of the diaphragm. It is probably possible in this way 
to practically eliminate the loss due to stage leakage. 

HAMILTON-HOLZWARTH TURBINES. 

A steam turbine called the Hamilton-Holzwarth is being 
developed by the Hooven, Owens, Rentschler Company of 
Hamilton, Ohio, which is a slight modification of the Rateau 
type. According to designs which have been published, this 
turbine is divided into two sections (high- and low-pressure) 
which are separated by a bearing. The principal difference be- 
tween this and the Rateau type is that the nozzles are arranged 
in complete rings around the circumference of the diaphragms in 
all the stages, instead of being grouped at the high-pressure end.* 
As in the Curtis turbine, the blades and nozzles increase in radial 
height gradually toward the low-pressure end. 

The number of stages is about the same as in a Rateau turbine 
for the same conditions of pressure, superheat, and vacuum, so 
that the nozzles are always designed to be non-expanding. This 
turbine has not been developed commercially,! so that it is not 
necessary to give other details. 

* This method of having complete admission around the blade wheel makes 
very small (short) ones for high-pressure sections with too much edging and radial 
leakage around the blades. Making so many more nozzles for high-pressure sec- 
tions is also more expensive. 

f One of these turbines was operated for a few days at the St. Louis exposition 
in 1904, but the author has heard of no other important installation. 



COMMERCIAL TYPES 257 

THE ZOELLY TURBINE. 

The Zoelly turbine is a modified form of the multi-stage 
impulse type. It has fewer stages (about 5 to 10), and is gen- 
erally a much simpler design than a Rateau turbine. It repre- 
sents a noteworthy attempt at increasing the steam velocities in 
the blades; but with it results the great disadvantage that the 
surfaces of the numerous large wheels and blades, many of 
which move at high speeds in steam of high pressure, produce 
excessive losses due to fluid friction. This fluid friction of disks 
and blades increases, of course, enormously as the speed and the 
pressure of the steam are increased. 

In a Zoelly steam turbine there are a number of single impulse 
wheels, each rotating in a separate chamber, the walls of which 
are formed by stationary flat disks to which the nozzles are 
attached. At the high-pressure end the nozzles occupy only a 
portion of the periphery; but the area covered gradually in- 
creases till at the low-pressure end practically the whole circum- 
ference is covered. When there are about ten stages the pressure 
in no stage drops to less than .58 of that in the preceding stage, 
so that non-expanding nozzles are used. The blades are dove- 
tailed as represented by Fig. 61, and there is no shroud ring. The 
tops of the blades are cut off parallel to the shaft, but at the roots 
they are made with a considerably greater height on the dis- 
charge side than on the entrance side. This is of course desir- 
able to allow for the loss of velocity in the blades, but it is stated 
that the height is made unusually large to cause the steam to flow 
smoothly through them without producing eddies.* In order 
to accommodate, in the different stages, the size of the nozzles to 
the expansion of the steam, the radial widths of the nozzle parts 
are gradually increased toward the low-pressure end. The most 
interesting part of the design of this turbine is, however, in the 
construction of the blade wheels to resist the stresses due to 
extraordinarily high peripheral speeds. As the blades for this 
turbine are made at present, they are much longer in comparison 
* It is probable that there is considerable expansion of the steam in such blades. 



258 THE STEAM TURBINE 

with the size of the wheel than in any other turbine; in fact, the 
length of the blade is sometimes nearly one-half the radius of the 
wheel. These long blades are tapered off toward the outer ends 
in order to make them of uniform strength. The disks are 
made of forged steel and the blades of nickel steel which resists 
erosive action very effectually. This simple construction of the 
wheels and blades makes a great saving in weight. The large 
radial divergence of these long blades makes possible the use of 
\ery small angles on the discharge side of the wheel. 

Turbines of this type intended for condensing service are 
usually made in two sections — each about 5 stages — placed 
far enough apart to permit a bearing to be located to support the 
turbine shaft at the middle. 

Zoelly deserves the distinction of being the first to adopt, in 
impulse turbines, the use of blades with unequal angles at the 
entrance and exit sides. Simplicity in the design of the working 
parts is the most striking feature of these turbines. 

There are a number of manufacturers of Zoelly turbines in 
Germany and France. It is stated a Zoelly turbine has been 
constructed at the Providence Engineering Works, Providence, 
R.I. 

Stage Leakage between the diaphragms and the shaft of 
steam turbines resembling the Curtis and Rateau types is easily 
measured by computing the flow through the nozzles for dis- 
charging steam that are located in the diaphragm through 
which leakage is to be determined. For this calculation by the 
use of the formula (4), page 39, only the area of the nozzles 
actually open and the initial pressure at these nozzles are required 
in most types of impulse turbines of a ew stages; that is, when 
the final or discharge pressure of the steam from the nozzles is 
less than .58 of the initial pressure of the steam entering the 
nozzles. The difference between the actual steam consumption, 
preferably measured by a surface condenser, and the calculated 
flow is the stage leakage. 






COMMERCIAL TYPES 259 

PELTON AND SIMILAR BUCKET WHEEL TURBINES. 

Impulse turbines with bucket wheels of the Pelton type have 
recently received a great deal of exploitation from inventors in 
America. This type has probably received so much attention 
because the Pelton water wheel, commonly known as the " hurdy- 
gurdy " wheel, has proved so efficient in American water power 
plants where a high head is available. 

Professors Rateau in Paris and Stumpf and Riedler in Berlin 
have done a great deal of experimental work on such turbines, 
but they have practically abandoned them for those with blade 
wheels of the common axial flow type. Rateau has now adopted 
his famous " multicellular " type, and Riedler is engaged in 
developing the Curtis turbine in Germany. 

Sturtevant Turbine. A steam turbine has been developed by 
the B. F. Sturtevant Company, Boston, Mass., from designs 
prepared by Mr. W. E. Snow, which in the general bucket 
arrangement is similar to the old Riedler-Stumpf type.* This 
turbine was developed primarily for driving blowers, but it is, 
of course, equally applicable for other purposes. It is notable 
particularly for its extreme simplicity and strength. 

Fig. 143 is a good illustration of this turbine, showing the 
buckets on the wheel and the segments on the inside of the 
casing including the nozzles and the stationary " reversing " 
buckets. Three, four, or five of the latter are cut into the seg- 
ment, following each nozzle, depending on the velocity of the 
steam. Fig. 144 shows more clearly the arrangement of the 

* The unique feature of the Riedler-Stumpf turbine was in the bucket wheels, 
of which the Sturtevant wheel in Fig. 143 is a good illustration except that there was 
usually a double row of buckets on the rim of each wheel. These wheels were 
patented by Prof. Stumpf and developed with the assistance of Prof. Riedler. The 
buckets were cut into the rim of the wheel by a milling machine, and were arranged 
to overlap each other like the shingles of a roof, instead of being placed one in front 
of another as in a Pelton water wheel. Unusual attention was given to balancing 
the wheels, which were in the form of flat disks. Stumpf states that these disks 
were balanced so accurately that the center of gravity came within .004 of the 
diameter from the geometric center. These disks were similar to the design in 
Fig. 216. It is stated that such wheels were designed for a factor of safety of 5 
at a rim speed of 1200 feet per second. 



260 



THE STEAM TURBINE 



nozzle with respect to the stationary buckets. The nozzle is 
the nearly square opening shown next to the first bucket, count- 
ing from the left. As shown here the steam flow will then be 




Fig. 143. Sturtevant Turbine with the Wheel Removed to the Side to Show the 
Arrangement of the Buckets. 

toward the right from the nozzle into the bucket opposite it on 
the wheel. From this moving bucket the steam will be diverted 
back into the stationary bucket next to the nozzle and the steam 




Fig. 144. Sturtevant Nozzle and Stationary Buckets, showing Flanged Connec- 
tion to the Steam Chest. 

path continues alternately through moving and stationary 
buckets until the last stationary bucket has been passed, when 
it will escape into the casing and into the exhaust pipe. The 
stationary bucket shown to the left of the nozzle is called a 



COMMERCIAL TYPES 



261 



" supplementary " bucket intended to utilize the velocity of 
the steam escaping over the top of the first moving bucket oppo- 
site the nozzle. Its function is to divert this steam leakage into 



the moving buckets. 




Fig. 145. Section of Sturtevant Turbine. 

Fig. 145 is a sectional view of the turbine. Hand wheels are 
shown on valves by which the flow of steam into the nozzles can 
be controlled. It is thus possible to close some of the nozzles 
on light loads and obtain nearly as good efficiency and steam 
consumption as at full load. The method is the same as ex- 
plained for De Laval turbines on page 144. The governor of the 
centrifugal throttling type is shown at the extreme right-hand 
end of the turbine shaft. It is one of the type with weights 
acting on knife edges, in principle somewhat like the De Laval 
governor (Figs. 160 and 161). Like other parts of this turbine 



2 6 2 THE STEAM TURBINE 

it is made as simple as possible, consisting of very few parts as 
shown in Fig. 147. 

The main bearings have solid linings of phosphor bronze. 
They are of the self-aligning, ring-oiling type. The weight on 
these bearings never exceeds 14 pounds per square inch of bear- 
ing surface. 

The speed of these turbines is from 1600 to 3000 revolutions 
per minute. These low speed limits compared with the speeds 
of single-stage De Laval turbines are made possible by the 




Fig. 147. The Parts of a Sturtevant Governor. 

application of the velocity stage principle in the use of the 
reversing buckets. 

Fig. 148 is an illustration of a Sturtevant turbine direct-con- 
nected to a ventilating fan or blower. The governor mechanism 
is at the left-hand end. Valves for closing nozzles to adjust the 
steam supply to the load, to get the best efficiency of the nozzles 
and blades, are shown clearly outside the casing. 

The deep base which the small diameter of this turbine neces- 
sitates, is utilized for steam chambers, to which the main admis- 
sion and exhaust steam piping is connected. Overhead pipes 
are in this way eliminated. 



COMMERCIAL TYPES 263 

The bucket wheel is a single forging of open hearth steel, and 
as the buckets are cut out of the solid metal, a wheel of great 
strength is secured. Blade breakage and "striking" are elimi- 
nated, because if the bucket wheel should get out of line and 
touch the casing on its sides, the result would be merely like the 
rubbing together of two steel plates, which would produce no 
serious injury. 




Fig. 148. Sturtevant Turbine Direct-connected to a Blower. 

This turbine was designed to require the minimum amount 
of attention and repairs. It is stated that it can be operated 
continuously under ordinary conditions with no more attention 
then the weekly filling of the oil wells in the main bearings. It 
is therefore particularly well suited for driving any type of 
auxiliary machinery, especially such as may be located in inac- 
cessible places. Such turbines make operating expense and 
depreciation low, and it is stated by some engineers that they 
have operated turbines of this type for five years at a time 
without any expense for repairs. 

Kerr Turbine. An impulse turbine of the Pelton type has 
been patented by Mr. C. V. Kerr and was formerly manufactured 
by the Kerr Turbine Company, Wellsville, N. Y. In this tur- 
bine Pelton double cup-shaped buckets are used into which jets 
of steam at high velocity are discharged from nozzles, located 



264 



THE STEAM TURBINE 



as in Fig. 149, around the periphery of the wheel. The inside 
surface of each bucket is formed of two intersecting surfaces of 




Fig. 149. Kerr Bucket Wheel and Nozzles. 




Fig. 150. Sectional View of the Kerr Turbine. 

revolution, approximately ellipsoidal, somewhat like the reflector 
of a locomotive headlight. 

A section of this turbine is shown in Fig. 150. In this 



COMMERCIAL TYPES 



265 



design there are five compartments or stages, each with a single 
bucket wheel. In the design of this turbine provision was made 
for its manufacture in standard "unit parts." In this sense the 
turbine casing shown here in section consists of steam and exhaust 
end castings and a number of nozzle diaphragms between the 
ends. In the chambers thus formed steel disks revolve, each 
having a row of buckets dovetailed in the rim. By this simple 
arrangement it is possible to build up turbines of any size for 
any pressure and vacuum by adding sections of nozzles as may 
be required. The nozzles are screwed into steel nozzle bodies, 
which are accurately set in place and riveted into the diaphragm 
castings. 

The governor is of the centrifugal type, consisting of weights 
moving on knife-edges. A section of the governor weights and 




Section A-A 



Fig. i si. Section of Kerr Governor Weights and Mounting. 



their mounting is illustrated in Fig. 151. The weights are sup- 
ported at three points. The hardened knife-edge at B is straight, 
and of sufficient width for the stresses on it. At 90 degrees on 
each side is a rolling contact at C. The curve at this point is 
such that the bearing between the weight and the cam collar is 
always on the line of centers. Pure rolling contact is thus 
secured, and the weight, without being fastened in position, is 
firmly driven by its triangular * support. The outward move- 
ment of the weights compresses the governor spring and oper- 

* The triangular or 3-point support is on B and on the two rolling surfaces at C 
on each side of B. 



266 



THE STEAM TURBINE 



ates, through lever connections, a balanced piston valve con- 
trolling the flow of steam. 

Fig. 152 shows a Rateau type of turbine as now made by the 
Kerr Turbine Co. Because of the simplicity of the design these 
turbines are particularly suitable for " isolated " lighting plants. 
These are sometimes provided with gears. 

Terry Steam Turbine. Like the Sturtevant turbine, the one 
invented by Mr. Edward C. Terry belongs to the Pel ton impulse 
type in which there are two or more velocity stages. Stationary 
reversing buckets are arranged in groups — one for each nozzle — 




Fig. 152. A 100-Kilowatt Kerr Turbine. 



around the interior of the casing. These bucket groups are shown 
in Fig. 153, where a Terry turbine is shown with the upper half 
of the casing raised for inspection. In this illustration there 
are four stationary buckets for each nozzle. Obviously the steam 
is returned to the moving buckets as many times as there are 
stationary buckets in each group. These stationary buckets are 
made of gun metal, and each has a crescent-shaped hole at the 
center through which the steam partially exhausts. There is, 
therefore, apparently considerable expansion in the moving 



COMMERCIAL TYPES 



267 



blades. A valve is provided for each nozzle, so that when it is 
desired some of them can be closed. Speeds of these turbines 




Fig. 153. Terry Turbine with the Casing Raised. 



vary from 2500 for a 10-horsepower size to 1600 for 300-horse- 
power. 

Fig. 154 is an illustration of a Terry turbine direct-connected to 
a five-stage high-pressure turbine pump. 



268 



THE STEAM TURBINE 




COMMERCIAL TYPES 



269 



Dake Steam Turbine. This turbine has stepped buckets, but 
the nozzles do not discharge radially. Stationary and moving 
buckets, with a section of t^e bucket wheel, are shown in Fig. 155. 



Steam Port 
Expansion "Wedge 
Nozzle 




Fig. 15: 



Steam Passages in a Dake 
Turbine. 



Fig. 156. Diagram of Nozzles and 
Buckets of a Dake Turbine Showing 
Expansion Wedges. 



This arrangement is unique in that the steam passes through the 
stationary buckets b in an axial direction, and is deflected radi- 
ally by the two sets of buckets a and c on the wheel. Steam 
enters the nozzles from the steam chest S. Relative positions of 
nozzles and buckets are illustrated diagrammatically in Fig. 156. 
Actually in the turbine the horizontal lines in this figure are, of 
course, arcs of circles. The " steps" shown here in the wall of 
the bucket ring are intended to bring the surfaces upon which 
the steam impinges nearer to the nozzles and to present always 
approximately the same angle to the flow of steam. 

The nozzles are designed with the object of delivering the steam 
to the buckets in parallel jets. Throughout their lengths the 
nozzle walls are the same distance apart, and expansion is secured 
by the use of " expansion wedges," shown plainly in both figures, 
which are set centrally in the nozzles. These wedges can be 
readily removed and replaced, so that it is not difficult to insert 
a wedge properly proportioned to give the best expansion for a 
given steam pressure. 



270 



THE STEAM TURBINE 




W 
>^ 

M 

O 



a 

GO 



COMMERCIAL TYPES 



271 



SPIRO TURBINES.* 
The Spiro turbine (Fig. 157) consists simply of two herringbone 
gear wheels which mesh together and revolve in a close-fitting 
casing. Steam enters at the inlet pipe at the bottom and passes 
around the gears in its expansion to the exhaust pipe at the top. 
Steam discharges from the inlet pipe through the small holes, 
equivalent to non-expanding nozzles shown in Figs. 158 and 159, 
into the " pockets " or spaces between adjacent gear teeth. As 




Fig. 158. The Spiro Casing or Cylinder. The two holes near the central rib 
inside the cylinder are the steam nozzles. 

the rotors revolve the " tooth-space " occupied by the steam 
increases in length as the steam expands. Finally the steam 
escapes when the outer ends of the teeth pass the line of contact 
between the two rotors. The increased length of this " tooth- 
space " from the time the steam is admitted until it is exhausted 
is shown in Fig. 160, by the comparison of the length of the 
tooth-grooves at " A " with the length of the outer white lines. 
By having the steam inlet at the bottom the weight of the rotors 
is partly carried by the steam pressure and friction is much 
reduced below what it would be if the inlet were at the top. 

* It is generally understood that a steam turbine is a machine for transforming 
the velocity of steam into work. The Spiro turbine operates only by the expansion 
of steam, and is therefore more correctly a rotary engine. 



2f2 



THE STEAM TURBINE 



Spiro turbines are suitable only for non-condensing operation 
and find application usually for driving the auxiliaries like 
blowers and pumps in a power plant or in office buildings where 
the exhaust steam is passed through feed- water heaters or is used 



EXHAUST 




INLET 



Fig. 159. Section of Spiro Cylinder and Rotors at Mid-length. 




Fig. 160. The Rotors. 



for heating buildings. Under these conditions low steam con- 
sumption, as would be obtainable with condensing operation, is 
unimportant. Sufficient expansion for condensing operation is 
impracticable with this type of turbine. Compactness is also 
an important feature. The casing of such a turbine is alone no 



COMMERCIAL TYPES 273 

larger than the cylinder of a good high-speed engine of the same 
capacity and is even smaller than comparable commercial sizes 
of electric motors. Governing is accomplished by throttling the 
steam pressure by a method similar to that used for nearly all 
steam turbines of relatively small size (less than 100 horsepower). 



CHAPTER VIII. 
GOVERNING STEAM TURBINES. 

Methods of governing steam turbines, or, in other words, of 
regulating the supply of steam to suit the load on the machine, 
may be classified as follows: 

i. Throttling or partly closing the steam admission valve. 

2. Varying the cross-section of the steam passages by " cutting 
out nozzles. " 

3. Varying the time of admission, or "blast" governing. 

4. Admitting steam at boiler pressure at various points along 
the direction of steam flow, or " by-pass" governing. 

When the steam admission of a turbine is partly closed, the 
amount of steam passing through the valve ports is, of course, 
reduced; but at the same time the steam is throttled, meaning 
that the pressure is reduced without changing the heat contents 
in a unit weight. Although in this throttling process the total 
heat in a pound of steam remains unchanged, the energy avail- 
able from expansion is considerably reduced. If steam at 165 
pounds per square inch absolute pressure which contained ini- 
tially 2 per cent, moisture is throttled without loss of heat, that 
is, without doing work, to 25 pounds per square inch absolute, 
the steam at this lower pressure will have 40 degrees F. of super- 
heat. Now if the available energy is calculated for adiabatic 
expansion from this lower pressure and 40 degrees F. super- 
heat to 1 pound absolute, it is found to be 207 B.T.U. The 
available energy for adiabatic expansion of steam at the initial 
condition before throttling (containing 2 per cent, moisture) 
to the same final pressure is, on the other hand, 316 B.T.U.* 

* Although the moisture is removed and the steam is superheated there is no 
gain to offset the loss in available energy except that the disk and blade rotation 
losses are reduced; but the gain from this cause could not probably exceed 10 per 
cent. " Drying " action then is not very important, and it will have very little 
influence in remedying large losses due to throttling. 

274 






GOVERNING STEAM TURBINES 



275 



In this extreme case of throttling, the available energy of the 
steam is reduced about 35 per cent, and consequently, for the 
same work, approximately 35 per cent, more steam is needed with 
throttling valves than if the steam could be used at light loads 
without throttling. It is not unusual for turbines governed by 
throttling to take steam at full load at 135 pounds pressure 
when the steam supplied is at 165 pounds. Then the 
maximum pressure becomes available only on overload just before 
the stage valves open. Efficiency of an expanding nozzle is 
considerably reduced when it is used with pressures very much 
different from that for which it was designed, as shown by the 
curve in Fig. 28. Blade efficiencies are similarly reduced when 
the available energies and consequently the velocities are not 
those for which the blades were designed. Fig. 161 shows very 



8 h - 





\ \ 














V 


A 












1 
















N^ 
























id 


2C 


30 


40 


5C 


K) 6C 






Load-Kw. 



Fig. 161. Effect of Throttling on Steam Consumption. 

plainly the effect of throttling on the economy of steam turbines. 
The two curves in this figure show the steam consumption per 
electrical kilowatt for a 600-kilowatt turbine when operating 
(1) with a throttling gpvernor, and (2) with a governor varying 
the steam supply by changing the area of the steam passages, that 
is, governing without appreciable throttling. In spite of these 
defects, however, governing by throttling has been fairly satis- 
factory. 

In the De Laval, Rateau, and Zoelly turbines governing is 
effected by throttling devices. For most turbines, the governor 
itself is similar to centrifugal governors used in reciprocating 
engine practice. 



276 



THE STEAM TURBINE 




Fig. 162. De Laval Governor and Vacuum Valve. 




FfG. 163. Section of the Main Admission Valve of a De Laval Turbine, 



GOVERNING STEAM TURBINES 277 

Fig. 162 shows cross-sections of a typical De Laval governor. 
It consists of two half cylinders B, B which are pivoted in a short 
outer casing by the knife-edge A. Inside the casing these cylin- 
ders are fitted with pins C, C which press on a collar D when the 
other ends of these cylinders (at B) are thrown out, or tend to 
separate, by centrifugal force. The pressure on the collar D 
transmitted by the pins compresses the springs and forces a 
central spindle G toward the right, which moves with it the 
bell-crank L. This bell-crank moves a short shaft which passes 
through the steam pipe and has attached to its other end by 
means of a set-screw a lever (shown in the section of the valve, 
Fig. 163) operating the main admission valve. The weight of 
the valve and levers is balanced by the small spring N. The 
bell-crank L has a certain "play" in M which is adjusted to 
make the governor not too sensitive to momentary changes in 
speed. The valve travel is only about one-eighth of an inch 
from the closed to the wide open position. 

The governor frame is supported on a tapering rod E which is 
fitted into the end of the main turbine shaft K. 

With condensing De Laval turbines a vacuum valve T is 
arranged in connection with the governor to act as an emergency 
stop valve. In case the turbine exceeds the allowable speed 
limit due to the failure of the main admission valve to operate 
properly, the vacuum valve admits air to the turbine exhaust 
pipe through the passage P. The steam consumption when 
operating non-condensing is so much greater than when condens- 
ing that it is said to be impossible to exceed the rated speed when 
exhausting into the atmosphere, with all the nozzles and the main 
admission valve open. 



Fig. 164. A Slide Valve Arrangement for a Turbine. 

Governing by " Cutting Out Nozzles." One of the simplest 
forms of governing is represented in Fig. 164 showing a plain 



278 THE STEAM TURBINE 

slide valve arrangement for regulating the flow of steam through 
a series of nozzles. This is one of the best systems that can be 
employed for an impulse turbine if an elaborate valve gear is to 
be avoided. 

As the full initial pressure is always maintained in all the nozzles 
that are open, there can be very little throttling except when the 
valve is in a position so that one of the nozzles is partly covered. 
The loss, however, due to this amount of throttling is practically 
negligible for other than very light loads. Valve gears have been 
designed to improve on this slide valve method by providing a 
separate valve (usually of the poppet type) for each nozzle or, 
at most, for a small group of nozzles. These valves are opened 
and closed suddenly by the governing apparatus by the use of 
either springs or dash-pots, very much as with our modern Cor- 
liss valve gears for reciprocating engines. The difficulty with 
this last method is, however, that there will be abrupt, although 
perhaps small, variations in speed every time a valve opens or 
closes, unless special precautions are taken in the design. If 
the service is for electric lighting, speed irregularities due to such 
governing may be sufficient to produce a flicker in the lights. 

In turbines with more than one pressure stage, as, for example, 
in the Curtis and Rateau types, it has often been proposed to 
control the admission to each stage. Apparently the only objec- 
tion to such a scheme would be in the very complicated valve 
gear that would be needed; but, contrary to what one might 
expect, it can be shown by tests and demonstrated mathemati- 
cally that such an arrangement would not give as good economy 
as if only the first stage is controlled. 

The only advantage resulting from this method of controlling 
the steam supply is that by making the light load pressures more 
nearly in the same proportion to each other as for full load and 
overload, the stresses in the diaphragms separating the stages are 
more nearly the same as calculated in the original design. There 
is probably no commercial type of turbine using such a compli- 
cated method of governing except for large overloads, when econ- 
omy is not of importance and the conditions are more of emergency 



GOVERNING STEAM TURBINES 279 

than of continuous operation. Governors for the larger sizes of 
the Curtis turbines show the merits of this method to the best 
advantage. By governing in this way it is possible to vary the 
number of valves supplying steam to the turbine in proportion 
to the size of the load, thus maintaining a constant initial pres- 
sure and therefore constant velocity in the nozzles and blades. 
In a single stage turbine there is no difficulty in applying this 
method, and consequently the energy and velocity are always 
those suited for the best nozzle and blade efficiency. Usually 
in a turbine of several stages no attempt is made to regulate 
the number of nozzles after the first stage, on account of the 
mechanical difficulties inseparable from a complicated valve 
gear. In order to secure a correct energy and velocity distri- 
bution throughout the turbine, the nozzles in all the different 
stages should, of course, be changed in the same ratio. This 
scheme is not impossible and has been attempted in some German 
designs. With turbines like the Rateau and Parsons, where the 
drop of pressure is very small in each stage, and where there 
are, therefore, a great many stages, any method of cutting out 
some of the steam passages to reduce the area at light loads is 
impracticable. 

Types of governors to be used depend a great deal on the 
capacity and the kind of service. The smaller sizes have usually 
simple forms, while the larger ones are necessarily more com- 
plicated. On the small turbines, where an elaborate valve gear 
is not desirable, the valves are moved by the direct action of the 
centrifugal force of the governor weights. This is called direct 
governing, to distinguish it from the " relay" system used by 
most turbine manufacturers for large machines. By the direct 
method a comparatively large centrifugal force is necessary to 
move the valves; and unless they are carefully balanced it is 
difficult to make the governor sensitive to fluctuations in the 
load. Besides, if for any reason a valve sticks, there may be 
wide variations in speed. 

By the indirect or what is commonly called the "relay" 
method the centrifugal force of the governor is needed only to 



28b 



THE STEAM TURBINE 




a, 



c 



X 



282 THE STEAM TURBINE 

"give the signal," as we may say, which sets in motion an auxil- 
iary mechanism by which the valves are moved by gearing con- 
nected to the main shaft or by steam or hydraulic pressure. In 
Curtis turbines of all sizes up to 500 kilowatts the valves are 
operated mechanically, and for larger sizes a hydraulic apparatus 
is used. 

Electromagnetic Control of Valves. Formerly, in large Curtis 
turbines, the valves opening the nozzles were operated by the 
pressure of steam admitted through a port opened and closed by 
a "pilot" valve controlled by electromagnets. The governor 
was connected to a very simple mechanism for the purpose of 
making and breaking the current through the electromagnets, 
which, in turn, moved the "pilot" valves operating the main 
valves on the turbine. 

Mechanical Valve Control. One of the recent developments 
in the valve gears for large turbines governed by cutting out 
nozzles is the successful replacing of the electromagnetic "relay" 
outfit formerly used on Curtis turbines by a positive mechanical 
valve gear, due to Mr. Richard H. Rice. 

This valve mechanism is well illustrated in Figs. 165 and 166, 
where it is shown applied for regulating the steam admission to 
the first stage of an impulse turbine. Steam in the steam chest 
C is maintained constant at the pressure for which the turbine 
was designed, and the valves are operated so that they are always 
wide open or else tightly closed. When the valve rod t, Fig. 166, 
is raised steam is admitted through the port A, from which it 
passes into a nozzle plate (like Fig. 114) at B to be discharged at 
high velocity into the blades of the first stage. 

The valve gear consists essentially, besides the worm gears 
shown at the right-hand side of the figures, of a connecting rod 
moving a bell-crank 1, to which two dogs or "catches," w, w, are 
attached by pins. The extreme ends of these dogs, marked 1 
will engage with the teeth on the steel plates u and v. An 
eccentric, h (Fig. 165), gives the connecting rod k a reciprocat- 
ing motion which, being transmitted to 1, moves the dogs w, w 
up and down. In Fig. 166 the lower dog is shown sliding on 



GOVERNING STEAM TURBINES 283 

the plate u, and in its lowest position it touches the tooth on 
this plate. The upper dog is kept out of contact with the tooth 
on the plate v by the lever x, which by engaging with the lower 
end of this dog, marked 2 in the figure, raises the end 1 out of 
reach of the tooth. The letters x and s are at opposite ends of 
the same lever supported on the shaft m. In the top view of 
this valve gear shown in Fig. 165 there are five valves operated 
by the connecting rod k. On the same eccentric, h, there is 
also another similar connecting rod, j, operating five valves on 
the opposite side of the turbine. The steam supply of the tur- 
bine is therefore regulated by ten valves. 

The position of the end of the lever at x is regulated by means 
of the rod q, which is connected to the Curtis governor illustrated 
in Fig. 167. Speed regulation by means of this governor is 
accomplished by the balance maintained between the centrifugal 
effort of moving weights and the static forces exerted by springs. 
The governor is keyed to the main turbine shaft at S and, of 
course, rotates with it. It is protected on two sides by a stationary 
looped casing, of which a section is shown at the top of the figure. 
In the order of action of this governor the weights A fly out on 
account of centrifugal force, moving on knife-edges near their 
largest diameter, and pull down the governor rod C by the pres- 
sure exerted on other smaller knife-edges B. The governor rod 
is pulled down against the action of the heavy spring D. At E 
a ball-bearing gimbel joint, thoroughly lubricated, forms a 
junction point between the revolving shaft of the turbine and 
the stationary lever of the governor (shown in the figure 
extending toward the right, nearly horizontally).* This 
stationary lever is connected by means of a bell-crank to the 
rod q (Fig. 166) and thus determines the position of the 
lever x. 

To illustrate the action of this valve gear and the governor, 
assume the load on the turbine has been increased and the speed 

* Connected to the stationary lever of the governor is an auxiliary spring F for 
varying the speed when synchronizing. By means of a small motor G the tension of 
this spring can be adjusted from the switchboard. 



284 



THE STEAM TURBINE 



has dropped a little, indicating that more steam is needed and 
that the valves have been so arranged * that the one shown in 
Fig. 1 66 is the next to be opened. With a reduced speed the 
governor weights A (Fig. 167) will move in slightly toward the 
center, reducing the tension on the governor spring D, so that 




Fig. 167. Sectional View of Curtis Governor. 

the rod C and the left-hand end of the stationary governor lever 
are raised. By means of an auxiliary lever and a bell-crank the 
rod q is raised and the end x of the lever attached to it is lowered 
to engage the catch 2 of the lower dog w, releasing at the same 
time the upper dog, which now comes into contact with the tooth 
in the plate v, raises the valve rod t, and admits steam through 
the port A. When again the speed becomes too high the rod q 
is lowered, x is raised, and the lower dog closes the valve. The 

* Each of the levers c controlling the dogs is set at a little different angle to the 
horizontal. The lever which has its end (x) lowest will open its valve first. 



GOVERNING STEAM TURBINES 285 

dogs are held in position when not in contact with the lever x 
by the flat springs o. 

The eccentric h is moved by means of gears connected to the 
main turbine shaft S. A ring on this shaft has a single tooth a, 
which engages with a gear wheel b, on the shaft c, which by 
means of worm gearing is connected with another horizontal 
shaft f, and thus moves the eccentric h. The hub of a turbine 
wheel is shown at H (Fig. 166), and carbon packing rings to pre- 
vent the leakage of steam from the first stage are illustrated by 
D and E. The speed reduction is designed sometimes for one 
worm gear instead of two as shown here. This is equivalent to 
putting the eccentric h directly on the shaft c. 

Hydraulic Motor Control of Valves. The hydraulic governing 
device used in the designs of Curtis turbines of the 500 to 9000 
kilowatts sizes is illustrated in the following figures. The move- 
ment of the horizontal governor lever shown in Fig. 167 is trans- 
mitted through the rod D (Fig. 168) to a second lever arm C 
operating the pilot valve P of the oil cylinder B. The piston A 
operates the main power arm of the mechanism, which trans- 
mits the motion either by a rack connecting with a pinion, or by 
means of cranks, to the " side " rod, shown in Fig. 169, carrying 
the cams for operating the valves. These cams act directly on 
the valves, opening and closing them according to the demands 
of the load. Because this device has a very slow motion it has 
the advantage of being practically independent of lubrication for 
its successful operation * 

Governing by Varying the Time of Admission. Governing 
by periodic admission or by " blasts " was invented by Parsons 
and has been applied to practically all types of steam turbines 
using his name. In its ideal form steam is admitted to the tur- 
bine by a poppet valve in puffs or blasts in periods of long or short 
duration depending on the demands of the load. The method 
is explained usually by saying that there are alternate periods 
when the turbine casing is either filled with steam or there is no 

* This pilot valve admits oil under pressure through the passages at ends of the 
"chest," and discharges oil through the exhaust port at the middle. The fluid 
being always under pressure makes movement slow and the amount of movement 
is proportional to the position of the pilot. 



286 



THE STEAM TURBINE 




, HVTTHT^fS 



f -<f~Y 



! rrn i 

ill 




r3~» H 



ittj.j LJaLl 



t iiMJ 



>P*n- 



tf 






i : 



Fig. i 68. The Hydraulic Operating Mechanism for Valves of a Curtis Turbine. 



steam at all. At light loads the valve opens for short periods, 
remaining closed the greater part of the time. When the load 
increases the valve remains open longer, and at about full load 



GOVERNING STEAM TURBINES 



287 



there is full pressure in the high-pressure blades, the valve 
merely vibrating without sensibly affecting the pressure of the 
steam in the passages. It is thought that in this way the full 
benefit of high-pressure steam can be secured at all loads. This is 
the ideal condition, but practical considerations greatly modify it. 




Fig. 169. One of the Two Valve Sets of a Curtis Turbine Showing the 
Hydraulic Controlling Cylinder and the Valve " Side " Rod. 

Governor for Curtis Poppet Valve Gear. The operating gov- 
ernor (Fig. 170a) is of the inertia type, mounted on the upper 
end of the vertical shaft which drives the gear pump. The 
combined rotary and vertical motion of the governor link is 
transformed into vertical motion by a ball bearing transmission. 
The cut shows the governor in its casing with the lid cut away 
to show the working parts. 

The motive power of the governor depends on the unbalanced 
force due to the inertia and centrifugal effect of the two weights 
(18) pivoted on ball bearings (17) opposed by the tension of a 



288 



THE STEAM TURBINE 



helical spring (3) which connects the two weights (18) through 
plugs (8), bolts (13) and knife edges (12) which bear on the knife 
edge seats (11) in the weights. The center of gravity of the 







Fig. 170a. Governor for Curtis Poppet Valve Gear. 

weights is inside the radius of the point of support and in advance 
of it with respect to rotation, thus decreasing the tendency to 
" jump " on quick changes of load and increasing its steadiness. 



GOVERNING STEAM TURBINES 289 

The bolts (13) can be screwed in and out of the plug for the 
purpose of obtaining different tension adjustments on the spring 
and are checked with lock nuts. The governor body is fastened 
to its shaft by a key and set screw, the latter being accessible 
through a hole in the governor casing. 

The governor is . permanently adjusted before leaving the 
factory. The connecting rod is trammed at both ends and all 
lock nuts locked with check washers. 

The speed regulation is changed as follows: In order to lower 
or raise the speed without materially affecting the variation of 
regulation, add or subtract weight. Changing the spring tension 
affects the regulation, narrowing upon increasing and broaden- 
ing upon decreasing tension. The change of speed in either case 
is about twice as rapid at the lower end of the travel as at the 
upper. The regulation can be further broadened by screwing 
the plugs into the spring and narrowed by unscrewing them, 
but this method should be resorted to only when the other two 
methods prove insufficient. 

By comparing the desired regulation with the actual, and 
applying the above methods in proper proportions, any desired 
regulation within the limit of design can be obtained. To illus- 
trate, assume the regulation to be high and narrow; in order to 
broaden, put in weight and slacken the tension on the spring 
at the same time. Should the regulation be high and broad, 
enough weight must be put in to bring the speed below [nor- 
mal and the speed raised again by increasing spring tension, etc. 
The method of adjusting length of the rods and levers leading 
from the governor to the valve mechanism is comparatively 
simple. 

Curtis Poppet Valve Mechanism. The mechanism for oper- 
ating the valve (Fig. 170b) consists of a chamber which includes 
the piston (33) and pilot valve (25); a controlling chamber or 
steam chest which includes the various valves (5) and valve 
centers (4); and finally the connections tying the governor to 
the operating and controlling chambers. In detail the operating 
cylinder carries piston (33) with its piston rings (28) . The piston 
is securely fastened to piston rod (32) which in turn is tied 
through coupling (37) to connection rod (38). The pilot valve 



290 



THE STEAM TURBINE 



(25) and its bushing (26) is carried in an extension of the operat- 
ing chamber. 

The controlling chamber or steam chest carries the valve stem 
(18) on which are mounted either four or five valve centers. 
Fig. 170b shows four valve centers, designed successively to raise 

£7 £3 29 00 3/ 3Z.&&3435 




6 5 4 3 Z / 



Fig. 170b. Curtis Poppet Valve Gear. 



r 



the various valves (5) with upward travel of the piston (33). 
The valves are of the poppet type, annular in shape, free to 
move but guided and controlled by the valve centers. The 



GOVERNING STEAM TURBINES 291 

valves together with the valve seats (6) are faced with a special 
valve metal. Leakage along the valve seats is prevented by 
the asbestos gaskets (7). The locking screw (14) acting on the 
steam strainer cage (9) holds the valve seats in place. 

Steam is admitted to the steam chest through the throttle 
valve not shown and passes through the strainer screen (8) . The 
opening of the various valves admits steam to the port plate, 
nozzles and turbine wheels. The operating chamber receives 
its steam through an opening in the casting connecting with 
bushing (26) and leading to the steam chamber formed by the 
pilot valve and its bushing. Ports connect this steam chamber 
with the piston chamber and steam is admitted to the upper or 
lower side of the piston, depending on the position of the pilot 
valve. The valve stem (18) and the piston rod (32) are connected 
to the governor through the pilot valve lever (29), connecting 
rod (38) and governor lever (39). 

With turbine and governor at rest and with no steam bled to 
the pilot valve, the position of the various levers is such that 
the pilot valve is in a position to admit steam to the under 
side ol the piston for opening the main valves. With the open- 
ing of the throttle, the turbine speeds up, the governor trans- 
mission moves upward and the levers move in a position to 
shift the admission of the steam from the under to the upper 
side ol the piston, closing the main valves successively until 
the governor takes a definite position. The floating lever (29) 
then fulcrums on the pin (42) and the pilot valve takes a neutral 
position independent of the governor. Any change in the posi- 
tion of the governor due, to a change in load or steam condition 
or both is immediately transmitted to the floating lever (29) 
which now fulcrums on the pin (30), and the pilot valve imme- 
diately moves to admit steam to the under side of the piston 
with a drop in speed, and to the upper side of the piston with 
an increase in speed. Sufficient steam will be admitted to the 
turbine to maintain a constant speed condition, at which time 
the governor again takes a definite position. 

It will be seen from the above description that throttling is 
practically eliminated since all but one of the valves is either in 
the wide open or closed position and the overtravel of the piston 



292 



THE STEAM TURBINE 



due to the expansive effect of the steam causes the opening and 
closing of the individual valve a sufficient number of times per 
minute practically to eliminate its throttling. A dashpot (48) 
filled with light cylinder oil prevents violent fluctuations. 

Brown-Boveri-Parsons Governing Device. The method of 
regulating the steam supply by intermittent admissions or 

" blasts " is typical 
of nearly all the gov- 
erning devices fitted 
to Parsons turbines. 
The design used for 
the Brown-Boveri- 
Parsons turbines is 
illustrated by Fig. 
171. Steam enters 
the turbine through 
a main admission 
valve N which is 
given a vertically os- 
cillating motion. A 
small piston mount- 
ed above this valve 
and on the same 
spindle has steam at the pressure in the main steam pipe on 
its lower face acting against the pressure of a strong spring 
on its upper face. An auxiliary valve fitted on the spindle L 
is given an oscillating motion by an eccentric on the governor 
shaft at M, which causes, at every stroke, the small passage 
at the lower face of the piston to communicate with the ex- 
haust, making the main valve N fall upon its seat. The 
spindle L is linked up to a collar sliding on the governor 
shaft. The height of the governor balls determines the posi- 
tion of this collar. Thus the height of the governor augments 
or diminishes the amplitude of the oscillations of the auxiliary 
valve on L, and in consequence causes the main valve N to open 
a longer or a shorter time at each admission of steam. The 
frequency of the steam admissions is about 150 to 250 per min- 
ute according to the speed of the turbine. 




Fig. 171. Brown-Boveri-Parsons Governing Device. 



GOVERNING STEAM TURBINES 



293 



Westinghouse-Parsons Governor and Valve Gear. Diagram- 
matically, the governor and valve gear of a Westinghouse- 
Parsons turbine are shown in Fig. 172. A small pilot valve, 
marked A in the figure, is actuated directly by the governor by 
means of levers and links. This pilot valve controls the steam 
supply of the turbine by regulating the operation of the main 
poppet admission valve which opens and closes at uniform 
intervals when the turbine is in operation. Speed variations 




Fig. 172. Diagrammatic Arrangement of the Governing Mechanism of a 
Westinghouse-Parsons Turbine. 



change the height of the governor balls which, in turn, change 
the position of the collar F of the lever on the governor spindle. 
By means of a system of links this lever varies the throw of the 
pilot valve relatively to the valve port. This pilot valve controls 
the main admission valve by means of the auxiliary piston valve B 
in the same way as in the Brown-Boveri design which has 
already been explained. Reciprocating motion for operating 
the valve mechanism originates in an eccentric driven by the 



294 THE STEAM TURBINE 

turbine from a worm on the main shaft. This eccentric gives an 
oscillating motion to the levers supported at D, F, and E. 

The governor is of the fly-ball type, the ball levers being 
mounted on knife-edges instead of pins, to secure sensitiveness. 
The speed of the turbine may be varied, while running, within 
the limits of the governor spring by grasping a knurled hand 
wheel at the top of the governor and bringing the spring and 
tension nuts to rest. Adjustment of the tension of the spring 
can then be made. This device is particularly useful for syn- 
chronizing the speed of small turbine-alternators operating in 
parallel, or for distributing the load between them. For syn- 
chronizing large Westinghouse turbine-alternator units a small 
motor controlled from the switchboard is used to adjust the 
governor spring. 

Allis-Chalmers Governor Mechanism. The governing device 
of the Allis-Chalmers steam turbine is of the Parsons type, using 
hydraulic instead of steam pressure. The governor is required 
to operate a small balanced oil relay valve only, while the two 
steam valves, main and by-pass, are controlled by oil pressures 
of about 20 pounds per square inch, acting upon a piston of 
suitable size.. The by-pass valve opens when the turbine is 
required to develop overload or the vacuum fails. 

The oil supply to the bearings and to the governor can be 
interconnected so that the governor will shut off the steam if the 
oil supply fails. 

"Blast" Governing Compared with Throttling. When the 
main steam admission valve of a Parsons turbine closes there is 
still some steam in the turbine casing, and this steam expands, 
of course, to fill the space. The same effect occurs also when 
the valves are first opened, and the steam rushes into a region 
of very low pressure. In these two ways low pressures are pro- 
duced just as with throttling valves, although, for the same 
average pressure, the loss is not nearly so great. The pressure 
variation for a 1500-kilowatt Parsons turbine at one-quarter, 
three-quarter, and full load is shown in Fig. 173. At a 
little overload there is practically no variation because the 



GOVERNING STEAM TURBINES 



295 



steam valve is then closed for shorter periods. Probably the 
greatest disadvantage from this method of governing results 
from "initial condensation at light loads." There is usually one 
blast or puff of steam in every thirty revolutions. Steam admis- 
sions are, therefore, far enough apart to allow the interior of 
the turbine to be cooled by the falling temperature between the 
blasts. Now when there is a fresh admission the steam comes 
into contact with the relatively cooler walls of the interior of the 




Atmospheric=0 

Absolute Zero 

Eodgkinson, F. 
Fig. 173. Indicator Cards Showing Initial Pressures in a Parsons Steam Turbine. 

turbine, and condensation must take place just as in a recipro- 
cating engine. 

If it were possible, practically, the number of "periods" would 
be made so small that free expansion would be reduced to a 
minimum; but for a satisfactory speed regulation long periods 
are not permissible. It appears, therefore, that unless the 
periodicity can be made low, the economy at light loads is no 
great improvement on the method of plain throttling. A very 
important feature of this method, however, should not be over- 
looked. This is the advantage of having a valve mechanism 
which is constantly moving, precluding the possibility of "sticky" 
valves. 



296 the steam turbine 

The time required for the steam entrapped in the casing when 
the valves are closed to drop in pressure by a given amount can 
be calculated very simply as follows. 
Let 

W x = weight of steam (pounds) entrapped when the valves 

are closed, 
W = weight of steam (pounds) in the turbine casing after 

expanding a time /, 
W f = weight of steam (pounds) flowing per second when 

the valves are wide open, that is, when the pressures 

in the casing are those for which the blading was 

designed, 
Pj, = initial absolute pressure of the steam delivered to the 

turbine, 
P 2 = final absolute pressure after a time /, 

and to avoid complex mathematical terms assume that in the 
expansion in the casing, in general terms, 

Pv =K, 

where v is the volume of a pound of steam and K is a constant. 
Then since v and W are reciprocals, 

W = P X constant; 
then also 

W = P x X C, 

where C is another constant. 
In a time dt we have thus 

dW = CPdt, 
also 

v 
W K ' 
K 



GOVERNING STEAM TURBINES 



297 



therefore 



W{u x dP 



\^i at, — - 

K 


> 


^T"* 




5? '-©='-■ 


-«» 


P^ = i£ and 




r IF' 






© 



If we take for convenience W = 2 J pounds of steam per second, 
W x = j pound, and P t = 165 pounds per square inch absolute 
pressure, the time required for the average pressure in the casing 

3 ,_ 165 

100 



to fall to 100 pounds is 



log e -77^ = .159 seconds. The 



4 X 2.25 

time required for steam at 165 pounds absolute to fall to various 
other pressures is shown in Fig. 174. 




Time-Seconds 

Fig. 174. Time Required for Pressure Variations in the Casing of a 
Parsons Turbine. 

With 165 pounds per square inch absolute initial pressure, 
usually the no load pressure varies from 25 to 50 pounds absolute, 



298 THE STEAM TURBINE 

when, according to the curve, the time required to reach this 
pressure without throttling is about .4 to .8 second; and as the 
load is increased correspondingly shorter times. 

Governing and Oiling Systems of Westinghouse Turbines. 
The valve control system, which might be spoken of as " oil 
governing," as now used on many Westinghouse turbines of 
over 1000 kw. capacity is shown in Fig. 174a. The governor is 
driven from the main shaft through a worm and gear. It con- 
trols and operates the primary and secondary steam valves. 
Oil for operating the relay is furnished by the turbine oil pump, 
which is situated below the goverr or and is driven through a set 
of gears on the end of the governor shaft. Two types of pumps 
are used, a simple rotary for the smaller turbines and, for the 
larger, a reciprocating pump driven by a Scotch yoke arrange- 
ment, transforming the rotary motion of the pump shaft to the 
reciprocating motion of the pump pistons. It was found expe- 
dient to use the latter type of pump when relatively large quan- 
tities of oil (as 80 gallons per minute) are circulated. The oil 
pressure on the discharge side of the pump is about 65 to 80 
pounds per square inch pressure (gauge), varying for different 
turbines. This high-pressure oil is used only in operating the 
oil relay, as the oiling system proper is supplied at about 4 to 8 
pounds per square inch pressure. This reduction in pressure is 
obtained through a reducing valve, which may be adjusted to 
give any discharge pressure desired. 

High-pressure oil is carried direct to the oil relay, A, shown in 
Fig. 174a. Depending upon the position of the oil relay plunger, 
this oil is admitted either below or above the oil operating 
cylinder piston, B, and so raises or lowers the oil operating pis- 
ton, closing or opening the valves and so admitting a smaller or 
larger amount of steam to the turbine. It may be noted that 
the secondary valve does not come into play until the lever arm 
of the operating cylinder piston connected to the secondary valve 
has moved until the adjustable screw on the top of the secondary 
valve comes in contact with the lever; so, as is obvious, only 
the primary valve is affected until the machine is loaded up to 
the point when the secondary valve comes into play. 

The action of the complete mechanism is as follows: Assume 



GOVERNING STEAM TURBINES 



299 



gmqanx "TO 







3 oo THE STEAM TURBINE 

the machine running at some load. If the tendency is for the 
load to increase, before the valve has had time to act, the speed 
will obviously fall off. This decrease in the speed causes the 
governor balls to close slightly, and through the linkage the float- 
ing lever is raised. The operating piston remaining stationary 
(for the moment), the oil relay plunger is raised, admitting oil 
above the operating piston and at the same time connecting the 
operating cylinder below the piston with the discharge outlet, 
which communicates with the general oiling system. Then the 
operating piston moves downward until the oil relay plunger 
shuts off the oil supply, at the same instant closing the oil dis- 
charge opening. If the operating cylinder has moved down- 
ward too far, admitting too much steam to the turbine, for the 
load required, the speed increases and the reverse action takes 
place; that is, the oil relay plunger moves downward, admitting 
oil to the lower side of the operating cylinder. 

Normal speed regulation of the turbines using this system is 
about 2 per cent, from full load to no load, and for instantaneous 
charge from no load to full load or vice versa the speed varia- 
tion will not be over 5 per cent., which shows the sensitiveness 
of the governor action and that of the complete system. 

There is a vibrator on the governor which imparts an oscil- 
lating motion to the linkage which is imparted or transmitted 
to the oil relay and oil operating cylinder and valves, insuring a 
free and sensitive action of the whole. A varying amount of oil 
passes through the oil relay, due to the movement of the operat- 
ing cylinder, but as the oil for the oiling system is throttled, 
doing no useful work, it matters little whether or not any or all 
of the oil passes through the relay. 

The oiling system as used with Westinghouse turbine gen- 
erator sets is simple. As mentioned before, the pressure on 
the system is just enough to insure a positive flow of oil to the 
parts to be lubricated. The oil, after it passes through the 
reducing valve and that which has passed through the relay 
system, flows to the cooler. The cooler consists of a chamber 
filled with tubes (generally of a copper composition) through 
which cold water circulates. By means of a series of baffle 
plates, the oil flows in an undulating path, coming in contact 
with the tubes several times. The warm oil meets the warm 



GOVERNING STEAM TURBINES 



301 



water tubes and the cold oil the cold water tubes, giving the 
most efficient cooling arrangement possible. 

From the cooler, the oil goes to four main points, the two 
turbine bearings and the two generator bearings. The supply to 
the turbine bearing at the governor end of the machine is the 
largest, as it also lubricates the thrust bearing, the governor and all 
parts in the governor case. After serving its purpose, this oil 
flows by gravity to the oil reservoir, and from there to the pump 




Fig. 174b. Jahns Governor. 

section, where it is again started on its way to the oil relay and 
through the oiling system. A vent is open to the atmosphere 
from the reservoir or the suction side of the pump to allow air to 
escape that may have come into the system. The oil cooler and 
the reservoir are fastened to the side of the bed plate of the tur- 
bine on the side opposite the valves and steam chest. When 
facing the turbine-generator set from the turbine end, the stand- 
ard arrangement is to have the valves and the steam chest on the 
right-hand side and the cooler and reservoir on the left. 

Jahns Turbine Governor. The speed regulating governor 
used on some De Laval turbines and some other makes is of the 
Jahns type (Fig. 174b). It is unique in that it is mounted in 



3° 2 



THE STEAM TURBINE 



such a way that the weights move on roller bearings on a hori- 
zontal spindle. The weight W acts only indirectly upon the 
governor springs A, A and in such a manner that the spring pres- 
sure is not transmitted directly to the governor mechanism. 
The movement of the sleeves B, B in the horizontal plane is 
guided in both horizontal and vertical directions by small roller 
bearings. The heavy stresses due to the centrifugal force of the 
governor weights during acceleration and retardation are not, 
therefore, brought upon pin joints or other mechanical devices of 
light construction. The movement of the sliding sleeves on the 

spindle S is effected by 
the bell cranks C, C 
which engage both 
the weights and the 
sleeves through con- 
necting arms support- 
ed likewise on roller 
bearings. 

Wilkinson Govern- 
ing Device. An in- 
teresting type of valve 
gear has been applied 
to Wilkinson turbines. 
Fig. 175. Wilkinson Valve Gear. The g ener al arrange- 

ment is illustrated in Fig. 175 showing governor, eccentrics, 
and a series of valve casings. One of these casings contains 
what is called a governor nozzle which is connected mechan- 
ically with the eccentrics. This is a form of auxiliary valve 
of which the function is not primarily to discharge steam into 
the turbine blades but to admit steam into or eject it from the 
other valve casings for the purpose of opening or closing them. 
This governor nozzle is illustrated in Fig. 176. The important 
feature of the governor nozzle is a cone-shaped piston at the 
lower end of the valve rod passing through the stuffing-box and 
connected to the eccentrics as shown in Fig. 175. A cone-shaped 
jet flows continuously over this cone. The central chamber of 
the governor nozzle as well as the spaces around it except a narrow 
annular passage communicating with a similar passage shown 
by a circular section in Fig. 176 at the left-hand side of the 





GOVERNING STEAM TURBINES 



303 



central chamber contains, in normal operation, steam at the 
"admission" or initial pressure. 

The valves in the other casings are operated by the force 
produced in this annular chamber by the injector or ejector action 
of the cone-shaped jet. Steam to be admitted to the annular 
passage must pass around the cone-shaped piston, and the 
position of this piston with respect to the annular passage deter- 
mines the effective pressure of the steam operating the admission 
valves. When the cone-shaped piston is in its lowest position 
the steam in passing around it to enter the turbine nozzles opposite, 
produces an ejector effect in the annular passage; but when the 
cone-shaped piston is at the other end of its stroke the steam 
produces an injector effect in the annular passage. When the 
injector effect predominates the pressure in the annular passage 
is greater than that in the steam chest, while with the ejector 
effect predominating this pressure is considerably less. 





Fig. 176. Governor Nozzle. 



Fig. 177. Admission Valve. 



One of the admission valves is shown in Fig. 177. A small 
passage of circular section shown here in the wall of the steam 
chest communicates with the governing valve or governor nozzle. 
This passage communicates with one side of the piston valve 
illustrated here. A spring is provided to keep the valve 
closed when the pressure in the passages communicating with 



304 THE STEAM TURBINE 

the governor piston is the same as in the steam chest. When 
the pressure in the passages is, however, less, corresponding to the 
ejector effect, by an amount greater than the tension in the 
spring (about 25 pounds) the valve is opened. With the injector 
effect, on the other hand, the valve will be closed. 

All the admission valves operate together, as the pressures are 
approximately the same in each of them. The governor valve 
oscillates 150 times per minute. The position of its cone-shaped 
valve with respect to the annular passage communicating with 
the admission valves is determined by the height of the governor 
weights and is adjusted by means of the levers shown in Fig. 
175. By this arrangement the duration of the ejector effect 
opening the valves is controlled by the speed. 

All the valves open a fixed number of times in a minute, but the 
duration of the period they are open varies with the load. 

Because the governor nozzle is always open to the steam chest, 
steam is never cut off from this nozzle, but with careful designing 
in proportioning the sizes of the nozzles this is no particular dis- 
advantage. 

In its action this valve gear is not unlike the usual Parsons 
governing device which has already been explained. It is likely 
that both are affected by quasi throttling at very light loads. To 
some extent the magnitude of this effect would probably be in 
proportion to the length of the casing. 

By-pass Governors. In all turbines the area of the steam 
passages increases in going from the high-pressure end to the 
exhaust. Consequently it is possible to pass a larger quantity 
of steam through a turbine for an overload, by admitting high- 
pressure steam into the middle stages in addition to the steam 
coming through the high-pressure nozzles. This is accomplished 
usually by the use of an auxiliary valve which opens slowly when 
an overload comes on the turbine and admits high-pressure steam 
directly into the low-pressure stages. As the steam entering 
through the by-pass valve acts on fewer rows of blades than the 
steam admitted under normal conditions, obviously, of course, the 
method is uneconomical and should, therefore, be used only for 



GOVERNING STEAM TURBINES 



305 



emergency loads. When a by-pass valve is used, the turbine is 
designed to be large enough to carry a little more than the normal 
full load, at which, of course, it is most economical; and for 
overloads it is expected that the efficiency will be considerably 
reduced. 

All the makers of Parsons and Rateau turbines use by-pass 
valves. The Westinghouse turbine has by-pass overload valves 
under the control of the governor, so that they open automatically 
when an overload comes on, but on most turbines by-pass valves 
are opened by hand. 

Overload economy is not usually of great importance, so that, 
practically, it is considered more feasible to use overload valves 
than to install additional turbines. In turbines of the Curtis 
type, which can be made to take a large overload with the addition 
of only a few extra nozzles without increasing the other dimensions, 
by-pass valves for overload have no advantages. In Parsons 
turbines, as anticipated in the design, there is usually a falling 
of! in speed when the overload valves open. 




Fig. 178. By-pass Valve Designed by Brown-Boveri & Co. 

A by-pass governor is shown in Fig. 178, which is a diagram- 
matic sketch showing the method of admitting high-pressure 
steam to the low-pressure stages of a Parsons turbine. This 
particular device is due to Brown-Boveri & Co. This design 
shows the by-pass method applied to an exceptionally well-made 



306 



THE STEAM TURBINE 



turbine much used in Europe. The by-pass ports open only at 
overload, and the speed is regulated for small fluctuations by 
the throttling method. In the figure the centrifugal governor is 
marked . 9, and operates by means of levers a balanced throttle 
valve. The by-pass valve 7, on the other hand, is operated by 
the pressure on the piston 10. Since this piston and the 
by-pass valve are on the same valve stem, they are raised or 
lowered together according as the pressure in the steam chest is 
high or low. With a high pressure the piston rises, lifting with 
it the valve 7, thus uncovering the ports, shown on one side of 
the turbine at 6, which admit steam through the pipes 3, 4, and 5 
to different parts of the turbine casing. Obviously there is a 
considerable change in the power developed immediately after 
the steam is admitted to one of the pipes 3, 4, or 5; and the 
consequent fluctuation in speed is taken care of by the throttling 
governor 9, or by an electrical solenoid governor indicated at 12. 




Fig. 179. By-pass Valve Arrangement for a Parsons Turbine. 

A simpler type of by-pass valve and governor arrangement is 
illustrated in Fig. 179. The by-pass valve is here directly under 
the control of the governor. The governor is marked 9 in the 
figure and operates a by-pass piston valve 7. The steam enters 
the turbine through the steam chest over the by-pass valve. 
When there is no overload on the turbine, the steam passes 
through the side port, which is shown open in the figure, to the 



GOVERNING STEAM TURBINES 307 

steam space 2 below, and from here it passes through the turbine. 
When, however, there is an overload, the by-pass valve 7 is raised 
by the governor and high-pressure steam is admitted through 
the pipe 3 to some lower pressure stages. With still more over- 
load the ports for the pipes 4 and 5 are also opened and high- 
pressure steam is admitted to stages intended for still lower 
pressures. 

When in order to make repairs to the condenser equipment, 
or for other reasons, it is necessary to run a Parsons turbine 
non-condensing the by-pass valve is opened and high-pressure 
steam is admitted to the intermediate stage of the turbine. By 
this method, because of the larger area of the passages, more 
steam can be used and the turbine is able to carry full load 
without a vacuum, although, of course, at a sacrifice of economy. 

If turbines are designed to take a large overload without a 
by-pass, the turbine must be of correspondingly greater capacity 
than the full rating indicates. The best economy of steam will 
then be at the highest output, and not quite so good at three- 
quarter load and full load. This is the usual practice in designing 
impulse turbines; but in the large sizes of Curtis turbines special 
overload valves are provided. 

Curtis Overload Valves. In Curtis turbines of four or five 
stages, especially in the larger sizes, automatic valves are pro- 
vided to open additional nozzles in the diaphragm between the 
first and second stages at times of overload. The usual designs 
of such valves are similar to the one shown in Fig. 180, which is 
arranged to operate when the pressure in the first stage — due 
to a large flow of steam — is larger than the normal. This 
design consists essentially of a piston valve fitted with a spring of 
sufficient strength to balance the unequal pressures on its faces 
for normal operation of the turbine. In the position shown in 
the figure the valve is closed; that is, no steam passes through 
it from the first stage to the nozzles discharging into the second 
stage. The pressure on its upper face is that in the first stage, 
while the pressure on the lower face is approximately that in 
the third stage. 



3 o8 



THE STEAM TURBINE 



f2r<s& SCage 




SecondS&Q&e 



7fifrcf66age 



Fig. 180. Curtis Overload " Stage " Valve. 



GOVERNING STEAM TURBINES 



309 



As the flow of steam increases due to increasing the load with 
a constant nozzle area the pressure will become greater in each 
of the stages; but obviously the pressure will be increased much 
more in the first stage than in the other stages, and at an overload 
the difference in pressure between the first and third stages 
becomes great enough to overcome the resistance of the spring 
on the overload or stage valve, so that it is forced from its seat. 
Steam then passes through a port communicating with an extra 
set of nozzles discharging into the second stage. 

These valves should be adjusted by varying the tension of the 
spring, so that they tend to open and close within a compara- 
tively small range of first-stage pressure. 

If the adjustment of such a valve has not been properly made 
and the valve remains open with a load fluctuating in a wide 
range between overload and considerably less than normal, the 
economy may be seriously impaired, or if one of these valves 
remains in a partly closed position so as to throttle the steam, the 
economy will be affected at all loads. Such valves to be efficient 
should open and close abruptly. 

Experimental Data Concerning Governing. Something should 
be said about the experimental results at hand concerning the 
different methods of governing. Curves illustrating the effects 
of throttling have been shown in Fig. 161, but a more satisfac- 
tory comparison can be made from the following table.* 



Curtis 

Curtis 

De Laval 

De Laval 

C. A. Parsons Company 

Rateau 

Westinghouse-Parsons. . . 
Westinghouse-Parsons. . . 
Zoellv. . . , 



Rated Full 
Load. 



Kw 



500 

600 

20 

200 



500 

400 

1250 

35° 



Fraction of Load. 



<6 



76. 
76. 
82. 
76. 
78. 
77- 
78. 



80.3 



Full. 



100 
100 
iOO 
100 
IOO 
IOO 
IOO 
IOO 
IOO 



* Mechanical Engineer, Jan. 20, 1906. 



310 THE STEAM TURBINE 

Fractions of load given at the top of each column refer to 
fractions of the most efficient load. Steam consumption at the 
different loads is expressed as a percentage compared with the 
steam consumption at the most economical load for each partic- 
ular machine. In other words, if the economy of any of these 
turbines were as good at half load as at full load we should have 
in the table under the column for one-half load 50 per cent, etc. 

Results in this table must be used guardedly and not confused 
with steam consumption. For example, the De Laval 200-kilo- 
watt turbine appears to such good advantage here because the 
system of governing used for these tests was nearly ideal. Full 
load steam consumption, on the other hand, was high com- 
pared with any other make of turbine in the list. The Zoelly, 
Rateau, and the 20-kilowatt De Laval turbines used simple 
throttling governors. 

In the effects of governing, steam turbines and reciprocating 
engines are different. The usual type of reciprocating engine 
has its speed lowered as the load is increased. Within limits, it 
happens also that the fall in speed is accompanied by a smaller 
requirement for steam. On the other hand, steam turbine 
governing operates in the reverse order, in that a fall of speed 
due to heavy load results in slightly increased " relative " speed 
of the steam in the blades, and consequently more steam will flow 
through the turbine. 

The Curtis and the 200-kilowatt De Laval were governed by 
varying the number of nozzles to suit the load. The original 
Parsons turbines used the " blast " governor. These data lead 
to the conclusion that the Curtis and the experimental De 
Laval (200-kilowatt) give the best results as regards the method 
of governing. 



CHAPTER IX. 
LOW-PRESSURE STEAM TURBINES. 

Early in the period of steam turbine development it became 
apparent that these new types of prime movers were capable of 
operating with ratios of expansion far beyond those economically 
possible with reciprocating engines. 

In the discussion of the effect of vacuum on steam consump- 
tion the good results obtained with turbines running at a high 
vacuum were clearly shown. With a high vacuum the heat 
efficiency of a reciprocating engine is not nearly so good as that 
of a turbine, because it is not desirable to make the engine cyl- 
inders large enough to handle economically the great volume of 
steam we have to deal with when the exhaust pressure is very 
low. For pressures slightly above atmospheric, however, a first- 
class, slow-speed reciprocating engine has a slight advantage 
over the turbine. We can see, then, that a combination of a 
non-condensing reciprocating engine with a condensing turbine, 
the latter taking exhaust from the former, might well be sug- 
gested. 

Fig. 181 shows graphically the volumes of the steam in each 
of the five stages of a Curtis turbine and illustrates how rapidly 
the volume increases at very low pressures. Reciprocating 
engines may be designed to operate with improved economy up 
to 25 or 26 inches vacuum; but this is about the limit. Steam 
turbines, on the other hand, will operate economically * with 
steam at the highest vacuum practically obtainable. 

The initial pressure of low-pressure steam turbines is usually 
that of the exhaust steam from non-condensing engines. With 

* It is not always commercially profitable to design a plant for operation at an 
extremely high vacuum, as the first cost of condensers and auxiliaries is usually a 
deciding factor. 

3" 



312 



THE STEAM TURBINE 



steam admitted to the engine at 200 pounds and exhausted from 
the turbine at 28 inches vacuum, theoretically there is no differ- 
ence in the total economy of a unit consisting of a reciprocating 





















' 
















~~ 


' 












<o 










- 
























' 




~ 
























~~ 








































N. 




























































































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<0 
















i 


















































































J 




■0 












































■> 
















~ 




































"' 






* 






































































































n 


















■ 










































1 


• 
























~ 
















^ 




















































































































s 


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- 






cs 



3 « 

't. .3 

U a 



o a 



a" 

CO 



fr 3^ 



> 
^ u 



o 9 



est 



10 ^ 10 



engine operating with an exhaust turbine taking steam within 
the limits from 7 pounds to 15 pounds per square inch absolute.* 
The curve in Fig. 181a shows the energy (B.T.U.) made available 

* Proc. Inst, of Naval Architects, 1908. 



LOW-PRESSURE STEAM TURBINES 



3*3 



by expansion from 26 inches vacuum to higher. The rapid in- 
crease above 27 inches is an important consideration and makes 



a 27 

1 



26 





































































































































































































































































































































































































y 


/ 


































































































































































































































,/ 


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10 



20 30 40 50 60 

in B.T.U. available between 26" and 29" vacuum. 



70 



Fig. i8ia. Curve Showing Enormous Energy Made Available for Work by 
Expansions to High Vacuums. Experimental Data. 

it very essential to get a very high vacuum in the exhaust on 
account of the enormous energy obtainable. 

Flattening out of the curve of the specific volumes of steam, 
Fig. i8ib, is also interesting in this connection. The increase in 
volume as the expansion is carried beyond 28 inches of vacuum 
is not generally realized. 

The following table prepared by the General Electric Company 
shows the large amount of work that is made available by the 
use of turbines in connection with existing non-condensing and 
condensing plants, the steam being delivered to the steam en- 
gine at 165 pounds per square inch absolute pressure. 

Owing to the rapid development of the turbine industry for 
high speed work and the close attention on the part of designers 
to this branch of turbine applications, the " combination " sys- 
tem of reciprocating engines and turbines was neglected. Only 
recently the advantages of this system have come to be generally 



314 



THE STEAM TURBINE 



recognized, and particularly in connection with marine propul- 
sion. Parsons has never advised the installation of an " all 
turbine " arrangement for ships designed for a speed of less than 





Atmos- 
pheric 
Pressure. 


Inches of Vacuum. 


Pressure of steam at tur-1 
bine admission valve 
in inches of vacuum. J 


O 


4 


8 


12 


16 


20 


24 


Per cent, gained over 
output of engine when 
worked with high vac- 
uum, the turbine ex- 
hausting to a vacuum 
of 28^ inches. 


26.I 


26.5 


26.8 


26.3 


25-3 


23.6 


20 



15 knots, and for moderate or slow speeds his designers have 
recommended the " combination " system. According to one 
of his designs for a cargo vessel intended for a speed of n J 









29 ■ , ■— •— 






5 *' 


S 28 ^ 


° z 


s 7 


i Z 


«• / 


9 / 

3 07 J 


S 27 T 


> 2 


r 


1 


26 I— 







100 



200 



500 



300 400 

Cubic Feet per Pound. 

Fig. 181b. Curve of Specific Volumes between 26 and 29 Inches of Vacuum. 

knots, if provided with a reciprocating engine discharging steam 
into the turbine at 7 pounds per square inch absolute pressure, 
the steam consumption was estimated to be 15 to 20 per cent, less 



LOW-PRESSURE STEAM TURBINES 315 

than that of an "all turbine" arrangement, or of triple expansion 
engines of the type usually fitted to this class of vessel. The 
" combination " system gives a vessel also greater maneuvering 
power than if driven only by turbines. 

On land low-pressure turbines have been installed principally 
in connection with rolling mill engines in steel works and wind- 
ing engines in mines. In both cases the engines are stopped 
or are running practically idle a large part of the time. These 
engines are usually reversing and are operated non-condensing. 
When a low-pressure steam turbine is installed to take the 
exhaust from such engines an equal amount of power can be 
obtained from the turbine (at 28 inches vacuum) as from the 
engine, thereby doubling the power of the plant without increas- 
ing the consumption of coal or the size of the boiler plant. 

Throughout the country there are a great many reciprocating 
engines exhausting into the atmosphere, and the exhaust from 
these engines is often wasted. There is no doubt that when these 
plants need increased capacity an installation of exhaust turbines 
will be profitable, even in most cases where there is no supply 
of water for condensing, and cooling towers must be erected. 
, There are also many power plants equipped with high grade 
compound reciprocating engines operating condensing which have 
a highel efficiency (not steam consumption) when operating non- 
condensing, or at a comparatively small vacuum, than when 
operating condensing. In such cases the installation of low- 
pressure turbines is probably always profitable. As an instance 
of the uses of exhaust steam turbines the following paragraph 
is quoted from a report prepared by a company manufactur- 
ing large sizes of both reciprocating steam engines and steam 
turbines. 

" A compound reciprocating engine with cylinder ratios of 
3.5 : 1, say of diameters 28 inches and 52 inches, with 150 pounds 
initial pressure, may be assumed to have 1000 kilowatts econom- 
ical capacity when running condensing and having a steam con- 
sumption of about 22 pounds per kilowatt-hour. This engine 
if operated non-condensing should have valve gears adjusted to 



316 THE STEAM TURBINE 

develop 1700 I.H.P., when it would consume about 20 pounds of 
steam per I.H.P. per hour. This gives 30,600 pounds steam 
available for the turbine, allowing 10 per cent, of moisture in the 
exhaust of the reciprocating engine. The total amount of steam 
passing the reciprocating engine, however, being 34,000 pounds, 
30,600 pounds would develop not less than 1073 brake horsepower 
in the turbine. Allowing 94 per cent, for the mechanical effi- 
ciency of the reciprocating engine, the combined horsepower 
developed would be 2673 brake horsepower and the steam con- 
sumption of the two units 12.7 pounds per brake horsepower, or 
18 pounds per kilowatt-hour, which is a remarkable performance 
for engines of such capacities operating without superheat. 
Compared with the performance of the reciprocating engine 
running condensing, this gives 75 per cent, increase of power 
and 18 per cent, saving of steam." 

Fig. 181c shows a very important installation of 5000-kilo- 
watt steam turbine-generators in combination with reciprocating 
engines of the same power rating. A low-pressure steam tur- 
bine has been installed to take the exhaust from each engine. 

Condensing engines when changed to non-condensing operation 
do not necessarily have their capacity in horsepower reduced 
because of the great increase of back pressure against which 
they must then operate. Such reduction would appear, however, 
on first thought to be the natural result; but, contrarily, the 
capacity of such an engine when changed to non-condensing 
operation may be unaltered or even in exceptional cases may be 
actually increased; particularly is this the case if the engine is 
one designed for a high expansion ratio. Under these conditions 
the high-pressure cylinder must have enough volume to pass 
the required amount of steam, without having the cut-off come 
so late as to sacrifice all opportunity to use the steam with a 
reasonably good expansion. There is an interesting reason for 
the capacity of many compound engines not being reduced when 
this change is made. In this adjustment the cut-off of the high- 
pressure cylinder has been shifted to make it late enough so that 
expansion in the low-pressure cylinder will not cause a loop in 



LOW-PRESSURE STEAM TURBINES 




Fig. i8ic. Steam Turbines taking the Exhaust from Large Reciprocating Engines. 



318 THE STEAM TURBINE 

the indicator diagram or to a final pressure in the low-pressure 
cylinder when its exhaust valve opens, which is lower than the 
average pressure in the exhaust line supplying the turbine, 
which is also the engine exhaust pipe. As the result of this 
adjustment of cut-off in the high-pressure cylinder the top part 
of the indicator cards taken from it will be observed to be much 
enlarged and of greater area than before, the increase being in 
some cases greater even than the area which is lost at the bottom 
of the low-pressure diagram by raising its exhaust pressure to 
about atmospheric. But there are also many compound engines 
operating condensing in which the release in the low-pressure 
cylinder occurs when the pressure is relatively high, possibly as 
high as atmospheric. Now the application of a low-pressure 
turbine to take the exhaust from this engine would have the 
effect of very materially reducing the capacity of the engine, as 
the benefits to be obtained to avoid the loop in the low-pressure 
diagram have been sacrificed in the original design of the engine 
and there is no chance to increase the area of the top of the 
diagram. 

Engines designed to operate both condensing and non-con- 
densing have generally the valves adjusted so that normally 
there will be a rather high back pressure at the point of release 
in the low-pressure cylinder, so that this " looping" of its indi- 
cator diagram will be avoided when running non-condensing or 
on light loads. When carrying full loads or an overload, on 
the other hand, the expansion will not be complete and a serious 
loss results in such engines designed for operation with loads 
varying considerably. In this case when the low-pressure tur- 
bine is applied there will be no occasion for the high back pressure 
at release, and the valve setting can be changed when running at 
full load to exhaust the steam from the low-pressure cylinder so 
that the exhaust pressure will be atmospheric or possibly a half 
pound or a pound above to assist in getting the steam readily 
through the exhaust ports of the engine. For this condition the 
blades of the low-pressure turbine taking this exhaust should be 
designed for initial pressure approximately atmospheric when 



LOW-PRESSURE STEAM TURBINES 319 

getting the amount of steam used by the engine at full load. 
Now when the cut-off is shifted back by the governor to keep the 
inlet valve of the engine open only half as long as at full load, 
only half as much steam will be delivered to the turbine and the 
absolute inlet pressure to it will also be reduced to half its 
former value. The final result is that the expansion of the steam 
in the low-pressure cylinder of the engine will be about to the 
inlet pressure of the turbine and there will be no appreciable 
" looping " of its indicator diagram, indicating that the condi- 
tions as regards effective expansion of the steam are excellent. 

In large power plant installations, it is the best practice to 
adhere to the " unit system " throughout; that is, in providing 
a separate low-pressure turbine for each engine and also a sep- 
arate condenser for each turbine. This method, although con- 
siderably more expensive than that of passing the exhaust steam 
from several engines to a single receiver supplying a relatively 
larger low-pressure turbine is, however, much to be desired as 
it gives so much greater flexibility in the operation of the plant 
and reduces very much the liability to an enforced shut-down 
of the plant due to condenser troubles. Maximum load on a 
low-pressure turbine connected to the engine as here described, 
that is, without having a governor on its steam supply pipe, is 
usually reached in conventional designs when the pressure at 
the inlet to the turbine is about twenty pounds per square inch 
absolute. A relief valve must always be provided in this low- 
pressure supply line which should open to let steam out to 
atmospheric exhaust when this pressure is exceeded. Steam 
discharged through the relief valve is an excess above what 
can be used in the turbine and is obviously wasted. 

All possible precautions should be taken to prevent the leak- 
age of air into all that part of the system operating at less than 
atmospheric pressure. This air is very detrimental to the proper 
action of a good condenser as it reduces the attainable vacuum 
and consequently renders impossible the full gains to be ex- 
pected from the installation of low-pressure turbines. Such 
leaks occur most generally in the joints of the exhaust piping of 



320 THE STEAM TURBINE 

both engine and turbine, also in imperfectly tight relief valves, 
as well as through the stuffing-boxes on the piston rods of the 
low-pressure cylinder of compound engines. To eliminate these 
difficulties the piping should be examined and tested frequently 
by applying a lighted taper to all questionable joints to observe 
whether the vacuum inside the piping tends to draw the flame 
toward it, as would occur if there were a leak. Another pre- 
caution often necessary is to put a special type of stuffing box on 
the piston rods of the engines, these boxes being supplied with 
steam at a pressure slightly above atmospheric so that air leakage 
inward is prevented. 

When several engines are connected up to supply exhaust to 
a single low-pressure turbine, it is always most desirable to turn 
over each engine for several revolutions with the piping connec- 
tions arranged so that the engine just starting will exhaust into 
the^ atmosphere. This should be done in order to avoid dis- 
charging the air in the engine cylinders into the low-pressure 
turbine and into the condenser. If this precaution is not taken 
when several engines are operating and another is started the 
effect in vacuum reduction will be observed. 

In matters of design, the low-pressure steam turbine presents 
no new problems. In fact its construction is in many respects 
simpler because of requiring fewer complicated details than the 
usual types of high-pressure or " complete expansion " turbines 
that have been studied. The relatively short length of the 
shaft or drum required for low-pressure turbines makes for 
rigidity and freedom from vibration stresses. The skill of the 
engineer comes into play in this new field almost entirely in the 
methods of application to conditions that a few years ago were 
not thought of. The primary consideration in practically all 
these applications is to utilize as much as possible of the avail- 
able exhaust steam about the plant, either in the low-pressure 
turbine or in some still more efficient method. While it is the 
object to show here the great advantages of this type of prime 
mover, yet it should be pointed out that there are often condi- 
tions arising in power plant practice when exhaust steam can 



LOW-PRESSURE STEAM TURBINES 321 

be used much more efficiently than in any known type of prime 
mover. The maximum thermal efficiency of a low-pressure tur- 
bine cannot well be made to greatly exceed 10 per cent., and even 
with this low efficiency when exhaust steam is a by-product, with 
no other available use, the addition of such a turbine to the plant 
will in such cases produce a great saving in coal bills. In other 
cases, however, where exhaust steam can be used advanta- 
geously for heating water, as for example in feed- water heaters, 
hot- water vats in manufacturing processes, or for heating build- 
ings, it would certainly be false economy to use the exhaust 
steam in a turbine and install low-pressure boilers for heating 
water. In the cases cited the thermal efficiency of the process 
of heating water with low-pressure steam is usually about 80 
per cent., which is to be compared with the 10 per cent, efficiency 
of the turbine. In this analysis it must not be overlooked, how- 
ever, that a steam power plant operating non-condensing can 
use only a very small percentage of the total amount of its 
exhaust steam for heating the feed water. In fact in the aver- 
age steam engine plant operating non-condensing only about 
one-sixth of the exhaust steam can be used for heating the feed 
water, the other five-sixths being discharged into the atmos- 
phere through the exhaust head. 

Of first importance among the general considerations affecting 
low-pressure turbine applications is the providing of adequate 
facilities for the removal of water and oil from the steam before 
it enters the turbine. Very wet steam can have no considerable 
deleterious effect on the turbine compared with the disastrous 
results often experienced in steam engine practice. It has the 
effect, however, of increasing enormously the fluid friction in the 
turbine blades and therefore of reducing the output and raising 
the steam consumption. Oil, when clean and pure, is not nec- 
essarily objectionable in the turbine and will pass through with- 
out accumulating, but in cases where boilers sometimes foam 
and discharge sulphates and carbonates with the steam, these 
will mix with the oil and form a gummy deposit on the blades. 
This deposit is not ordinarily removed by erosion, particularly 



322 



THE STEAM TURBINE 



when steam velocities are not over 400 feet per second, and will 
often choke the steam passages between the blades. 

Until recently in America exhaust steam turbines were usually 
arranged to take steam directly from the exhaust pipe of the 
engines without intervening valves or governing mechanisms. 
A generator direct connected to the turbine will operate very 
satisfactorily with generators adapted for connection in parallel 
to engine-driven generators, and the turbine set thus "floats 
on the system." As it receives only steam exhausted from the 
engine its output will therefore vary as the load on the engine. 
When the load becomes light the steam supply will be reduced 
by the governor on the reciprocating engine. In the case of direct 
current units the generators may have shunt windings, and as the 
voltage will vary nearly as the speed, the load will be automati- 
cally proportioned between the reciprocating and turbine units. 




Fig. i8id. Simplest Combination of Low-pressure Steam Turbine and Recip- 
rocating Steam Engine. 

The most common and probably also the simplest application 
of the low-pressure turbine is shown in Fig. i8id. As shown the 



LOW-PRESSURE STEAM TURBINES 323 

generators connected to both the reciprocating engine and the 
turbine are connected to the same three-phase alternating-current 
circuit and here also no governor is used on the turbine. By 
this arrangement the turbine will take automatically its share 
of the total electrical load in proportion to the amount of steam 
supplied to it. If it tends to forge ahead of the reciprocating 
unit it will take more of the load, leaving less for the engine 
whose speed will immediately increase until its governor re- 
duces the flow of steam to both the reciprocating engine and to 
the turbine, thus controlling with one governor the amount of 
steam supplied to the complete system. In case the generators 
driven by the engine and turbine are of the direct current type, 
as the turbine forges ahead, and takes more load, the increase 
in its speed raises the voltage slightly, which puts more current 
through the fields of the generators and tends to reduce the 
speed. Self-regulation is thus admirably accomplished. Obvi- 
ously for the same reason it is possible to vary the speed of the 
turbine slightly by adjusting the field rheostat. 

The quantity of steam used determines obviously the relative 
amount of load carried by the low-pressure turbine; the greater 
the amount of steam the greater the proportion of load taken 
by the turbine, which is due, of course, to the variation of pres- 
sure in the receiver. As this pressure increases the total range 
of pressure available for the turbine increases, the heat available 
per pound of steam increases, and consequently more work is 
done, assuming, of course, a constant vacuum in the turbine 
exhaust. 

In the method of low-pressure turbine installation described 
in the preceding paragraphs, where the turbine operated with- 
out a governor of its own, the electrical machines driven (gen- 
erators) were of similar types; that is, all the current generated 
was supplied to a single line. It is not infrequent, however, for 
low-pressure steam turbines to be installed to operate with re- 
ciprocating engines in power houses where the generators on the 
engines are to supply direct current lines and the generators on 
the turbines are to supply alternating current for transmission 



3 2 4 



THE STEAM TURBINE 



to a distance. Such an arrangement is shown in Fig. i8ie. 
Obviously the engine and the turbine must each have its own 
governor. If the low-pressure turbine were arranged to take 
only the load from the alternating current line there would be 
much steam wasted when the direct current load happened to be 
heavy and the other light; and conversely it might be neces- 
sary to sometimes supply the low-pressure turbine with high- 




Fig. i8ie. Application of a Rotary Converter in the Combination of Low- 
pressure Turbine with Reciprocating Engine. 

pressure steam when the load on the engine was light. A most 
satisfactory method to avoid these difficulties is to install a 
rotary converter or motor-generator set as shown. Any in- 
equality of the two loads will then be taken care of and the load 
coming on the engine and on the turbine will be divided auto- 
matically to give the best results. This sort of arrangement 
might not be very satisfactory for taking care of electric lighting 
loads if there were likely to be exceptionally frequent reversals 
of the operation of the converter from alternating to direct 
current and vice versa, as there might be a voltage change of 
several per cent, which would perceptibly affect illumination 
until again adjusted at the switchboard. In most cases, how- 



LOW-PRESSURE STEAM TURBINES 



3 2 5 



ever, the demand for one kind of current will always predomi- 
nate, so that this sort of reversal is not likely to be troublesome. 
Another application of a low-pressure turbine is illustrated in 
Fig. 1 8 if, where the steam engine drives a line of shafting through 
a belt drive and the low-pressure steam exhausted from the 
engine goes to a turbine generator unit supplying an electrical 
transmission line. In this case the engine and the turbine must 
each have a governor, as the loads on the two machines are en- 



A.C.Line 




Fig. i8if. Application of Condenser in Combination System. 

tirely unrelated. The device adopted in this case to make the 
plant as economical of steam as possible is to operate the engine 
in such a way that the excess of steam, above that required for 
the low-pressure turbine, is discharged directly into the condenser. 
By this method the engine will operate at times at a fairly good 
vacuum as determined by the relative amounts of steam and 
absolute pressures in lines A and B. The governor on the tur- 
bine operates only the by-pass valve V, regulating the flow of 



326 



THE STEAM TURBINE 



steam from the engine exhaust into the condenser. When the 
load on the turbine is light this valve will be nearly wide open, 
deflecting only a small amount of steam into the turbine; but 
when the turbine load gets near its full capacity, the condenser 
by-pass valve will be nearly closed. It becomes thus possible 
for the engine to obtain the advantages of nearly full vacuum 
when the turbine is running light. It is a good practice to put 
the usual type of valve on the inlet pipe to the turbine which will 
also be controlled by the governor to prevent the turbine run- 
ning away on very light load. 

A very interesting type of installation is shown in Fig. i8ig. 




From. Turbine 



Fig. i8ig. Application of Synchronous Motor in Combination System. 

The method illustrated here consists in the installation of an 
electric motor of the synchronous type supplied with current 
from the generator driven by the turbine and having the pulley 
on its shaft belted to the line shafting driven by the engine. In 
case the electrical load on the turbine-generator set becomes too 
large for it to handle it will slow down slightly with the result 
that the synchronous motor will be driven from the line shafting 



LOW-PRESSURE STEAM TURBINES 327 

as a generator to supply more current to the electrical supply 
lines. The additional load coming on the engine as the result 
of driving the motor will cause the governor to open the inlet 
valve wider on the engine and admit a larger amount of steam 
to the system. Conversely, when the line shafting is overloaded 
the governor on the engine admits more steam to the system, in 
greater amount, however, than the turbine requires. This re- 
sults in a speeding up of the turbine and a forging ahead of the 
synchronous motor so that it acts now purely as a motor to 
assist in driving the shafting. 

Provisions for Intermittent Supply of Steam. An ingenious 
development, largely due to Professor Rateau, has been applied 
to cases where the supply of exhaust steam is intermittent, as 
in the case of rolling mill and winding engines. Rateau's device, 
called an accumulator, is used to bridge over the "dead periods," 
and by providing sufficient capacity it can be made to provide a 
practically constant supply for an exhaust turbine. 

Rateau's Accumulator. This regenerator or accumulator is 
shown in Fig. 182, illustrating longitudinal and transverse sec- 
tions. This regenerator consists of a large cylindrical shell partly 
rilled with water. When the engine exhausting into it is running 
the steam is delivered as a spray through the small holes in a num- 
ber of pipes immersed in the water. By this method some of the 
steam is condensed and gives up heat to the most of the water. 

As these accumulators operate usually with steam at atmos- 
pheric pressure, the entering steam will have a temperature of 212 
degrees F. and will tend to heat the water to that temperature. If, 
now, the engine stops, the supply of exhaust steam is discontinued, 
and the flow of steam to the turbine will tend to make the pres- 
sure fall off slightly so that 212 degrees F. will then be slightly 
above the temperature of boiling water at this lower pressure. In 
this way the water will be evaporated to supply steam as a boiler 
would. If, now, the engine starts again, steam will be delivered 
to the accumulator at a temperature slightly above that to which 
the water has fallen, due to the cooling effect of the evaporation 
for supplying the turbine, and the mass of water will again absorb 



328 



THE STEAM TURBINE 




LOW-PRESSURE STEAM TURBINES 329 

heat from the exhaust steam. Water has a higher specific heat 
than any other substance except hydrogen, so that it is a most 
suitable and convenient substance for heat accumulation. 

In actual practice it is more convenient to run the regenerator 
at a pound or two pressure above the atmosphere, as in this case 
the piping is not under vacuum, so that so much care does not 
have to be exercised to avoid air leaks. In certain cases, how- 
ever, it is desirable to run below atmospheric pressure. In this 
way the power of the primary engine may be augmented by 
letting it operate at a partial vacuum. Plants are actually run- 
ning with a delivery pressure to the turbines as low as six pounds 
below atmospheric pressure. 

On account of heat radiation from the accumulator, water 
gradually accumulates in it; but by means of a float trap shown 
at the right-hand side of the longitudinal section this excess of 
water is removed. 

If for any reason the engine shuts down for a considerable 
period, the supply of heat stored in the accumulator will become 
exhausted and the pressure will fall below the practical limit for 
operation of the turbine. To provide for such an emergency an 
automatic reducing valve is inserted in the piping to deliver live 
steam to the accumulator. There is also a relief valve on the 
accumulator through which excess steam will pass off into the 
air when the pressure becomes 3 or 4 pounds above atmospheric. 

The pressure in the accumulator should be always about five 
to ten pounds per square inch, gage pressure, or at least a few 
pounds above the atmospheric to avoid the possibility of air 
leaking into the system. 

The important consideration in the selection and designing of 
an accumulator for a low-pressure turbine is the length of time 
the regenerator will be expected to carry full load on the turbine 
without receiving any low-pressure steam from the engine or 
engines. Obviously the longer this time is, the greater the 
capacity required of the accumulator. Quite generally the mis- 
take has been made, according to Hodgkinson,* of supplying 

* The Electric Journal, April, 1913, page 335. 



330 THE STEAM TURBINE 

these accumulators in much too large sizes for the require- 
ments. In many cases the time interval has been assumed to 
be six to seven minutes, during which the accumulator must 
supply the steam, while more careful study shows that five to 
six seconds would have been a much better estimate. The case 
of a steel mill is cited. If the exhaust is to be taken from a 
blooming mill the time element should bear some relation to the 
period between the passes of an ingot, as well as to the maximum 
time from the last pass of one ingot to the first pass of the next 
ingot. The accumulator should not be designed, therefore, to 
cover such delays as would arise from the clogging of the mills 
or because a new ingot might not be ready to be bloomed. For 
these cases of unusual delays another method is recommended. 
When the demand for steam at the engines is interrupted there 
will be a sudden rise of pressure in the boilers and the safety- 
valves will blow off. This steam should be piped to the accu- 
mulator inlet instead of being allowed to escape to the atmos- 
phere. This steam from the safety valves will assist materially 
in helping the low-pressure turbine in carrying its load. A very 
good arrangement for the accomplishment of this idea is to place 
a " cross-connection " of piping between the steam main sup- 
plying the engine and the engine exhaust line, and to put into 
this line a globe type of spring loaded valve set to permit steam 
to pass through the cross-connection when the pressure is a few 
pounds lower than that at which the boiler safety valves will 
blow. 

For engine power plants, where the supply of exhaust steam 
is often stopped for long periods, the accumulator installation 
is usually dispensed with, and the low-pressure turbines are 
provided with piping to take steam directly from the boilers, 
in addition to the exhaust steam piping. (See pages 335 to 

338-) 

An exhaust steam turbine has, of course, relatively few rows of 
blades compared with ordinary high-pressure turbines. 

From several tests made with 500-kilowatt exhaust turbines 
in England, a steam consumption of 34 pounds per kilowatt-hour 



LOW-PRESSURE STEAM TURBINES 



33 1 



was obtained with 15 pounds per square inch admission pressure 
and 28 inches vacuum. 

The curve in Fig. 183 shows the steam consumption in pounds 
per horsepower-hour at the switchboard of a 500-kilowatt 
exhaust steam turbine of the Rateau type. 



1 1 1 II 1 II 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 II ] 1 1 


— -- - - 1 II II 1 1 II 1 1 11 11 1 1 1 1 1 1 1 u 1 1 11 11 1 














::::::: ::: s: _r__ . _:_.___ __- _ 


__• _T ___ __ -_4- - -- - ,. 


___ -*- 


S ZZ - --± - 


sa ■ ::: _y : _ : _ --Z-t - - ::::::_ :: 


- x 






* :::: _ _ _\ _ . _ __ ______ 


u _-_i : _ _ \_ . _ __ ___ ____ 


a :: : \ _ _ _ _ _ :::::_ 


g :: :>, 


m :: — : — : :: — s ::::::_::_ _ __ : : ::::: : 


•r : _t__ __ __, _____ __::__: 


, — _ _>__ _ _-_ — 


__ ::_ _ _ __ __ v__ _ :_: : 




a; 4 "— _ _ : _._ _ _ __ "__"___""_:: 


u s s _ : : ? ::::::: 


9 .... . _______ _.S„_ _ 3 " 


& ::::;__ ::: :::::::::::;:::: ::____; _ _ ___: ::: ::_":~ 








3 _ _ _ _ _ _ _ :_:::_: 


p ___::::: _ :_ _:::: _ — :::::: ::: ::::::::: 




w ::::"::_t_":::: : — _:: : _ : :: — ::~::~_~::_ 
















U 1 1 1 1 II 1 11 Mil II 1 1 11, 1 1 ML. 1 T Tl hi H 



300 



400 



500 600 

H.P.et Switchboard 



700 



800 



900 



Fig. 183. Curve of Steam Consumption of a Rateau Low-Pressure Turbine. 

Some tests quoted by Francis Hodgkinson on a Westinghouse 
low-pressure turbine made recently gave the following results: 



Steam Pressure, 
Lbs. per Square 
Inch Absolute, 
Dry and Satu- 
rated Steam. 


Vacuum in Exhaust, 
Inches Mercury Re- 
ferred to 30 Inch 
Barometer. 


Load in 
Brake Horse- 
power. 


Total Steam 
per Hour. 


Steam Con- 
sumption 
Brake Horse- 
power Hour. 


17.4 
12.4 


25.98 
2 5-99 


920 

472 


25,670 
17,487 


27.9 
37-i 


11. 8 

7-7 
• 5-2 


26.97 
27.03 
26.98 


59 2 
321 
102 


17,720 

11,980 

6,57o 


29.9 
37-3 
64.4 


n. 6 

8-7 


27.8 
28.00 


586 

458 


16,400 
13,920 


28.0 
3° -.4 


6.1 

4-5 


27.90 
27.99 


234 
114 


9,036 
6,248 


38.6 

54-8 



332 



THE STEAM TURBINE 



Fig. 184 is a copy of a shop drawing of a 1000-kilowatt Westing- 
house double-flow low-pressure turbine. The exhaust steam from 
the engines enters through the annular space H and is distributed 




Fig. 184. iooo-kilowatt-Westinghouse Low-Pressure Turbine. 



to the right and left sections of Parsons blading. The upper half 
of the drawing is a section of the rotor and shows the method of 
construction. The exhaust is discharged through the base as 
indicated by arrows. The openings I, I are provided for con- 
venient inspection of the blading. They are covered with suitable 
covers in which automatic relief valves are fitted. 

Economy curves of this turbine are shown in Fig. 185. The 
pressure of the steam delivered to the turbine was approximately 
atmospheric. The vacuum, as shown by the curves, was 27 \ 
inches for one test and 28 inches for the other. 

Another Westinghouse turbine built to operate in connection 
with high-pressure reciprocating engines gave the following 
results in a shop test: 

Initial steam pressure, 15 pounds per square inch absolute. 

Superheat, 40 degrees F. 

Vacuum referred to 30 inch barometer, 23 inches. 

Load, 1500 brake horsepower. 

Steam per brake horsepower hour, 35.5 pounds. 



LOW-PRESSURE STEAM TURBINES 



333 



In all these tests the exhaust was condensed in a surface con- 
denser, which assures accuracy in measuring the steam con- 
sumption. 

A Curtis exhaust steam turbine installed in Philadelphia 
receives the exhaust steam from reciprocating engines at a pres- 
sure of 15 to 1 6 pounds per square inch absolute, and exhausts 
into a condenser with an average vacuum of 28 inches. The 
turbine has no governor, but takes all the steam the engines 
will supply. The output over and above that obtained from 




Fig. 185. Curves of Steam Consumption of a iooo-kilowatt Westinghouse 
Low-Pressure Turbine. 



the reciprocating engines is increased about 66 per cent, for the 
same steam consumption. With dry, saturated steam at atmos- 
pheric pressure delivered to the turbine, the guaranteed steam 
consumption is 36 pounds per kilowatt at full load, and 40 
pounds per kilowatt at half load. It is stated that the actual 
test results are probably at least 10 per cent, better. 

Since the exhaust from reciprocating engines is often very wet, 
it is good practice to insert a steam separator in the steam pipe 
leading to the low-pressure turbine. 

Most applications of low-pressure turbines have been made in 
collieries and steel mills, where non- condensing engines are the 
rule. Results are, however, so satisfactory that the design of 



334 THE STEAM TURBINE 

new plants having a compound engine with a smaller low-pressure 
cylinder than is usually provided, which is to discharge its 
exhaust into a steam turbine is likely to become common, as 
giving better steam economy than can be obtained from either 
reciprocating engines alone or turbines alone. 

Low-Pressure Steam Turbines Combined with Gas Engines, 
There is also another field open to the low-pressure steam turbine. 
The hot cooling water from the jackets of large gas engines could 
be heated by the exhaust gases, and the low-pressure steam thus 
formed would drive a steam turbine. 



CHAPTER X. 
MIXED-PRESSURE TURBINES. 

It sometimes happens that there is an available source of low- 
pressure steam which it is desired to utilize for the development 
of power, but where unfortunately there is not always at hand 
a sufficiently large amount of this low-pressure steam to take 
care of the power requirement. To suit this condition it is not 
infrequent to provide in a steam turbine a high-pressure section 
to take steam at boiler pressure and thus help out the low- 
pressure section when its normal supply of steam is low. In this 
type of construction when, however, the supply of low-pressure 
steam is sufficient for the power requirements, the supply of 
high-pressure steam is cut off entirely by the governor; and on 
the other hand when the supply of low-pressure steam becomes 
again sufficient in quantity, no more high-pressure steam is 
used. 

The generous use of live steam in low-pressure steam turbines 
is not by any means as poor engineering practice as at first 
thought it appears. The obvious reason for admitting live steam 
to the turbine is that the supply of low-pressure steam from 
the engines is insufficient for the turbine requirements, and that 
consequently some of the engines have been relieved of their 
load more or less suddenly. The boiler plant continues, however, 
to make steam at the former rate, and the safety valves will soon 
blow of! unless the excess steam can be used in the power plant. 
By taking this excess of steam to the turbine to help in carrying 
its load, which we shall assume has not been reduced, will serve 
to use this excess of steam to the best possible advantage. It 
is not an unusual practice even to pipe the discharge from the 
safety valves on the boilers into the receiver in the low-pressure 
piping supplying the turbine. These conditions are met most 

335 



336 THE STEAM TURBINE * 

frequently in the suddenly variable loads in rolling mills and in 
hoisting operations where the reciprocating engine drives the 
rolls or hoists and the low-pressure turbine supplies a more or 
less constant electrical load for both lighting and comparatively 
light power requirements. If the intervals requiring the use 
of high-pressure steam are relatively long, as for example five 
to ten hours on the average, then a so-called "mixed turbine" 
type should be used. The accumulator method is also adap- 
table for the longer period but is expensive as regards first cost. 
The so-called "mixed" steam turbine has been a development 
of the applications of steam turbines to suit two important con- 
ditions of operation which are as follows: (i) the case where 
a low-pressure turbine is to be used to develop an amount of 
power for which there is not constantly available a sufficient 
amount of low-pressure steam to carry the average load; and 
(2) when there are large enough quantities of low-pressure steam 
at certain times to carry the load but at more or less long inter- 
vals there is no exhaust steam supplied at all. Both of these 
cases require the supplying of large quantities of steam from 
sources independent on the exhaust lines and live steam direct 
from the boilers is invariably the substitute. For this sort of 
service with widely varying steam pressures the mixed-pressure 
turbine has found acceptable application. In speaking of a 
mixed-pressure turbine in this chapter we shall think of one hav- 
ing separate high- and low-pressure portions in a single casing. 
A good example is shown in Fig. 185a, where the high-pressure 
portion is provided with an impulse wheel which is made easily 
removable so that when for long periods high-pressure steam 
is not needed it can be taken off. High-pressure steam enters 
at the steam chest opposite the nozzles in the "impulse" section. 
Low-pressure steam enters only the reaction blading through the 
vertical pipe coming up behind the "reaction" section. Such a 
turbine differs essentially from the ordinary low-pressure tur- 
bine which is provided with its own governor only in having 
under the control of the governor a special set of valves arranged 
to supply live steam to nozzles directing steam into a section of 



MIXED-PRESSURE TURBINES 



337 



high-pressure blades before discharging through the low-pressure 
sections along with a supply of low-pressure steam with which 
it mixes. A very common type of mixed-pressure turbine con- 




H.-P. Steam Chest 



Fig. 185a. Mixed-pressure Turbine. 

sists of an impulse wheel in which the energy drop from boiler 
to about atmospheric pressure is absorbed and the remainder 
of the energy is taken out by expansion in low-pressure reaction 
blading. An installation of this kind is illustrated by Fig. 185b. 
The valves under the control of the governor are adjusted so 
that no high-pressure steam is admitted until the valves on the 
low-pressure line are wide open. There are often excellent oppor- 
tunities for the installation of mixed-pressure turbines in con- 
junction with accumulators; but in every case a check valve 
must be provided between the accumulator and the turbine in 
the low-pressure line to prevent live steam getting back into 
the shell of the accumulator, which will probably not be strong 
enough to withstand the excessive stresses that might be pro- 
duced. 

In many mixed-pressure turbines the high-pressure section, 
if of a simple disk construction, is frequently made removable 



33& 



THE STEAM TURBINE 



as in Fig. 185a, so that the "windage" loss due to its revolution 
when not in use can be eliminated. A good estimate is that 
about 2 per cent, of the power of the turbine is lost in the air 
resistance of such a high-pressure section. 




Fig. 185b. Low-pressure Turbine with Live Steam Valve Installed at Peace 

Dale, R. I. 



Mixed-pressure turbines are not often used in sizes larger 
than about 2000 kilowatts. To meet the requirements of the 
larger capacities it is best to use a regular or " complete expan- 
sion" turbine in combination with a smaller simple low-pressure 
turbine. 



CHAPTER XII. 
MARINE TURBINES. 

One of the most important fields for the steam turbine is the 
propulsion of ships. In the mercantile marine the progress of 
the turbine had been extremely rapid, the first mercantile vessel 
propelled by turbines having been built only a very few years 
ago. That vessel had about 700 tons displacement, and developed 
3500 indicated horsepower, comparing with a tonnage of 45,000 
and 70,000 horsepower in the Lusitania and Mauretania. Care- 
ful trials had shown that at all speeds above 14 knots the turbine 
was more economical than the reciprocating engine, being 1 5 per 
cent, better at 18 knots, 31 per cent, better at 20 J knots, and 36 
per cent, better at 20.1 knots. In the Dover-Calais service it had 
been found that the turbine boats carried passengers at two knots 
greater speed with 25 per cento less coal per passenger than boats 
propelled with reciprocating engines. A saving in coal of about 
9 per cent, was computed for the turbine steamers belonging to the 
Midland Railway (England), as compared with similar steamers 
of the same company equipped with reciprocating engines. 
The difference in initial cost and in weight of machinery was 
found to favor the turbine driven ships by ij and 6 per cent, 
respectively. When used for marine service, doubtless the 
greatest defect of practical steam turbines is that they cannot be 
reversed. Many attempts have been made to devise a turbine to 
reverse in a simple way comparable with a reversing reciprocat- 
ing engine. It is the present practice to provide turbine driven 
ships with two turbines used only for reversing, and as they are not 
intended for high speed, they may be of small power compared 
with the main turbines. These two reversing turbines are usually 
fitted to the same shafts as the low-pressure turbines, and when 
the ship is running ahead their rotors revolve idly in a vacuum. 

345 



346 THE STEAM TURBINE 

When the ship is to be run backward the steam is shut off from 
the "ahead" turbines and is admitted to the auxiliary reversing 
turbines. There is, of course, a disadvantage from not having 
at times the full normal motive power of the ship available 
for backing. Besides, conditions are not ideal when a large 
portion of the plant is idle for a greater part of the time. These 
reversing turbines will occupy a great deal of longitudinal space, 
so that the floor space required for an installation of marine 
steam turbines is larger than that required for reciprocating 
engines for the same conditions of service. 

The White Star Company (International Mercantile Marine 
Company) has decided to operate ocean steamers with a com- 
bined reciprocating and turbine engine plant. The two outer 
shafts will be driven by quadruple expansion reciprocating en- 
gines and the central shaft by a low-pressure turbine operated 
by the exhaust steam from the low-pressure cylinder of the 
reciprocating engines. For going backward, the reciprocating 
engines will be used, as they are readily reversed, and in the 
ordinary service the turbine and reciprocating engines will be 
operated together. By this combination the advantages of 
reciprocating engines for reversing are secured, together with 
the great range of expansion which is possible with the steam 
turbine. 

It is difficult to say what developments the future will bring 
in the applications of steam and gas turbines to the marine 
service. Practically all the new battleships and cruisers for the 
British navy are now turbine driven. If we consider that the 
steam turbine in its practical form commenced its real develop- 
ment only in 1885, the future certainly may have rich possibilities. 

Fig. i85i represents the results of tests made at variable 
speeds and powers on a standard combined impulse and reaction 
type of turbine. In explaining the results of these tests Mr. 
H. T. Herr * states that investigations now under way by the 
" Westinghouse interests " will insure the elimination in the near 
future of the reciprocating engine in the field of marine propul- 
* Journal of the Franklin Institute, March, 1913. 



BLEEDER OR EXTRACTION TURBINES 



341 



in the stage; those not throttled would be either closed off 
entirely or else wide open. 




Fig. 185CL Ring Valve of Curtis Bleeder Turbine. 




Fig. 185c Side of Diaphragm of Bleeder Turbine. 

Parts of this device are illustrated by the following figures: 
Fig. 185c! shows the ring valve used for covering the nozzles 



342 



THE STEAM TURBINE 



which are bolted in the usual construction to the side of the 
diaphragm (Fig. i8se). This valve is operated by the piston in 
an oil (or steam) cylinder which is in turn moved by being 
subjected to oil or (steam) underpressure admitted from a high- 
pressure supply by a small pilot valve actuated by leverage con- 
nections to the diaphragm (30) in communication by small piping 
with the stage in which the constant pressure is to be main- 
tained. Fig. i85f shows a cross-sectional view of the mechan- 




Fig. i85f. Valve Gear of Curtis Bleeder Turbine. 



ism which actuates the valve. By means of a flexible joint at 
(1) the piston rod (2) moves the valve plate back and forth over 
the face of the nozzles. Pressure on the piston (14) in the oil 
(or steam) cylinder (13) gives the movement to the piston rod. 
Movements of the piston are effected by means of the pilot valve 
(22) which is in turn actuated by the diaphragm (30) by means 
of the rods (34) and (39) . This diaphragm with its " corrugated " 
or " accordion " sides forms a cylindrical chamber which is in 
communication by means of small piping with the stage to be con- 
trolled, and from which the " bleeder " steam is to be taken. 



BLEEDER OR EXTRACTION TURBINES 



343 



Movements of the diaphragm are opposed by the spiral spring 
(36) which can be set to maintain any desired steam pressure in 
the stage. 

Careful inspection of Fig. i8sd shows that the ports in the 
ring valve are not all of the same size but are of progressively 
increasing width around the circumference from the narrowest to 
the largest. The narrow parts begin closing up on the first move- 
ment of the valve. There are four groups. The second group 
begins closing only after the first or narrowest set is fully closed. 

A balance-plate (Fig. i8sg) is put on top of the ring valve 
(Fig. i85d) for the purpose of assisting in equalizing the pres- 
sure on the two sides of the valve, and thus reducing the force 
required to move it, since the steam pressure is not effective 
on the whole surface of the ring valve. 




Fig. i85g. Nozzle-plate for Curtis Ring Valve. 

It is comparatively a very easy matter to remove some of the 
steam which has been partly expanded in the turbine by the 
use of suitable automatic or hand-controlled valves even when 
the quantity of steam required at a constant pressure in such a 
bleeder line is quite variable. By this method it is possible to 



344 



THE STEAM TURBINE 



extract the greatest amount normally possible as required for 
generating power and at the same time supplying at a reason- 
ably constant pressure " usually about atmospheric," or about 
5 pounds above, the requirements for heating or industrial 
purposes. 

Fig. 185I1 shows the satisfactory filling of the blades of an 
impulse turbine of the " bleeder " type. 




Fig. 185I1. Flow Lines in a " Bleeder " Impulse Turbine. 



CHAPTER XI. 

BLEEDER OR EXTRACTION TURBINES. 

The name " bleeder" or "extraction" turbine is given to 
one specially designed to take steam at boiler pressure and to 
exhaust part of this steam at a normally low vacuum while 
another part is " extracted " or taken out from one of the 
stages at a pressure of five to ten pounds per square inch gage 
pressure; that is, just a little above atmospheric. In many 
cases, as for example in cotton, woolen, and paper mills, this 
steam is "extracted" for manufacturing purposes, usually heat- 
ing water in vats. More commonly, however, such turbines 
find their application for supplying the low-pressure steam re- 
quired in a heating system for houses, factories, office buildings, 
etc. Because this latter supply is needed only a part of the year 
and otherwise is variable with the seasons, there will be times 
when the turbine operates by complete expansion of all the steam 
supplied to it by the boilers. Obviously it is necessary to provide 
in such turbines a means whereby the " bleeder " steam can be 
taken out at any time, and with sufficient back pressure even at 
light loads to maintain a pressure in the section from which the 
steam is to be withdrawn to overcome the resistances of pipes 
and valves, so that steam can flow freely as required. It is 
desirable also that the pressure in this section of the turbine 
should be fairly constant. To accomplish this result in Wes- 
tinghouse-Parsons turbines a partition diaphragm has been used 
to separate completely the high-pressure from the low-pressure 
portion, as shown in Fig. 185c. The steam enters the high-pres- 
sure section through admission valve A and passes normally out 
through the bleeder passage and into the mains to be supplied 
with steam. When the turbine uses more steam than is needed 
for the service supplied the pressure in these passages " backs 

339 



34Q 



THE STEAM TURBINE 



up," and the weight loaded by-pass* valve V opens when the 
pressure exceeds that for which the valve has been set. The 
steam then flows out into the low-pressure sections of the tur- 
bine and thence into the condenser. It is the function of this 
valve to maintain sufficient pressure in the passages to create 
the desired flow into the mains. There is usually some throt- 
tling action in valves of this type which may cause a slight drop 
in the available energy of the steam supplied to the low-pressure 
sections. 




Fig. 185c. Westinghouse-Parsons " Bleeder " Turbine. 



To avoid the throttling action, Curtis " bleeder " turbines 
are designed to regulate the flow of steam by the use of 
ring-shaped valve over the nozzles leading from the stage from 
which the steam is to be " extracted." This valve is operated 
automatically by a mechanism responsive to the pressure in the 
stage so that the effective area of the nozzles is changed as 
required to maintain a constant pressure. By this method the 
closing of the nozzles occurs only in groups, so that any slight 
throttling action that might occur due to partial opening would 
create its loss only in a small group rather than in all the nozzles 

* Similar to a relief or safety valve in its action. 



MARINE TURBINES 



347 



sion as the turbine generator has practically eliminated it for 
electric power plant service. 

On account of the difficulty of adjusting the inherent re- 
quirements of the steam turbine for operation at relatively high 
rotative speed and the corresponding difficulty, opposite how- 
ever, in effect, of the efficient operation of the propellers of 
steamships at anything but relatively low speed, the applica- 
tions of steam turbines to the propulsion of ships has been very 
much limited. If it were not for these difficulties there is no 
reason why the steam turbine should not displace the recipro- 



50 



^45 

a 

03 

§ 40 

a5 O 

■8*85 

■a -2 

M §30 

a * - 

(_ c 25 

a 1 20 



10 



2.5 



5 



^ 



aX 



-~-~J-- 



m 



70 g 



50': 



20 

18 o 

u 

<p 
16 £ 
o 
a 
<o 
14 g 
o 

12 | 

10 -° 



40 •§ 

m 

30 W 



Fig. 1851. 



800 1600 2400 3200 4000 

Revolutions per minute. 

Curves Showing Variation of Steam Consumption, Horsepower, and 
Efficiency of Latest Designs of Steam Turbines with Speed. 



eating steam engine almost entirely for this service. On this 
account, however, the application has been confined almost 
entirely to merchant and naval vessels designed for high-speed 
service. Experience has shown that in the applications of steam 
turbines to slow-speed ships there has been no appreciable 
saving in weight, in space, or in the cost of operation, over what 
it would have been with reciprocating engines. The nearest 
approach to the solution of this problem is to be secured prob- 
ably by the application of gearing essentially similar to that 
designed by De Laval. This sort of gearing cannot, however, 



348 THE STEAM TURBINE 

be applied without modification to turbines developing more 
than possibly iooo horsepower. A design much better suited to 
high-speed conditions and also adaptable for large power has 
been developed by the Westinghouse Machine Company with 
the cooperation of Mr. George Westinghouse, Admiral George 
Mellville, and Mr. John H. Macalpine. The essential principle 
embodied in their improvements consists in the application of a 
so-called " floating frame " designed to carry the pinion on the 
main turbine shaft. The experimental gear developed in the 
early stages had a floating frame supported on pivots, permitting 
flexibility as regards horizontal movement of the pinion, but 
was rigid as regards vertical movements. Very recently Mr. 
Westinghouse developed a very important improvement con- 
sisting in the substitution of a flexible support by means of 
hydraulic pistons taking the place of the rigid vertical supports, 
and in this way improving very much the efficiency and the 
wearing properties of the gear. This improvement in wearing 
properties had also the effect of reducing to a minimum all 
noise and vibration which in the original design were consid- 
erable. In this later design (Figs. 185J and k), the main frame 
supporting the pinion is held up by the pistons in the hydraulic 
cylinders filled with oil under pressure. This construction 
permits vertical movements of the pinion along with its flexi- 
bility in its floating frame and is therefore a great improve- 
ment in that the earlier design permitted only lateral move- 
ment. The vertical movement is permitted by the supporting 
of the floating frame on the piston connected to the support- 
ing rods. Similarly the lateral movement is permitted by the 
flexibility of the horizontal pistons on the two sides of the 
pinion. 

For a more complete description and discussion of the " float- 
ing frame " type of reduction gear see Engineering, vol. 95 (1913), 
pages 169 and 609, and The Electric Journal, January, 191 2. 

Water rate curves drawn from the data of acceptance tests 
of the battleships North Dakota and Delaware are shown in 
Fig. 185I. Curves A and Bi show the steam consumption per 



MARINE TURBINES 



349 




Fig. 185J. "Floating Frame" Reduction Gear, Showing Gears when Side of 

Casing is Removed. 



35° 



THE STEAM TURBINE 



_ i 


yf9 


B^» - 


if \$ 




Eg? « BB». ■*»— X 


JH L |\ 


L^Z^f 


• vV 

1 v 


MB 1 * 






jj» 1 m ^ iu *J 


1 





Fig. 185k. Reduction Gear, Showing Flexible Support of Pinion. 



22 
a 20 












































N 


1204 


KNO 


T8 

































Si 18 






S 


V 


































J? 16 




> 


«> . 


V 


^ 


«»» 






























14 

P4 


1 12 

1 10 

1 8 
6 


12-24 


knq; 


S» 




-<2l 








19-21 


CN01 


S. 










21 JS 


KNO - 


6 








- 














19-21 


fNOl 


s 












21-6 


UN 


nrs 





































































































































































2000 6000 10000 14000 18000 22000 26000 30000 
Horsepower. 

Fig. 185I. Water Rate Curves of U. S. Battleships and Computed Curves if 
Geared Steam Turbines were used. 



MARINE TURBINES 351 

shaft horsepower per hour of the engines in the two battleships, 
while C and Ci show the corresponding results if geared turbines 
had been used instead of reciprocating engines. It is estimated 
that with a geared turbine combination of the Westinghouse 
" floating frame " type the economy of the prime movers in 
vessels of the Delaware class could be improved 30 per cent, 
at full speed and 25 per cent, at cruising speed. Fig. 185m 




FlG. 185m. Low-pressure Turbine Casing for German Steamship. 

shows the enormous size of the casings for the low-pressure 
sections of the steam turbines installed in modern battle- 
ships. 

Electrical Transmission for Ships. Another method differ- 
ent from the use of reduction gears has been frequently sug- 
gested for making the steam turbine more adaptable for marine 
service. This method consists in using on the vessel steam 
turbines direct connected to high-speed electric generators, 
which can operate then under practically identical conditions 
as in " land " service. These turbines, obviously, can then be 
designed to operate at a speed best suited to obtain high effi- 
ciency. The electric current from the generators is used to. 
drive slow-speed electric motors on the shafts of the propellers, 
the speed here being that giving best efficiency for the pro- 
pellers. 

This method offers great flexibility in the handling of a 
vessel, as the motors can be very quickly reversed, and changes 
of speed are readily obtainable. 



352 THE STEAM TURBINE 

Although this method for marine propulsion has been ad- 
vocated by engineers for many years it has not as yet re- 
ceived very favorable acceptance; but, doubtless, it is con- 
stantly receiving more favorable attention from well-known 
designers. 



CHAPTER XIII. 

TESTS OF STEAM TURBINES. 

Testing Steam Turbines.* In every power plant the means 
should always be available for making tests of the steam equip- 
ment to determine the steam consumption. Usually tests are 
made to determine how nearly the performance of a turbine . 
approaches the conditions for which it was designed. The 
results obtained from tests of a turbine are to show usually the 
steam consumption required to develop a unit of power in a unit 
of time, as, for example, a horsepower-hour or a kilowatt-hour. 

In such tests a number of observations must be made regarding 
the condition of the steam in its passage through the turbine and 
of the performance of the turbine as a machine. To get a good 
idea of what these observations mean, it may be profitable to 
follow the steam as it passes through the turbine. The steam 
comes from the boilers through the main steam pipe and the 
valves of the turbine to the nozzles or stationary blades as the 
case may be. It then passes through the blades and finally 
escapes through the exhaust pipe to the condenser. It is pref- 
erable to have a surface condenser for tests so that the exhaust 
steam can be weighed. The weighing is done usually in large" 
tanks mounted on platform scales. 

Methods for Testing. The important observations to be made 
in steam turbine tests are: 

i. Pressure of the steam supplied to the turbine. 

2. Speed of rotation of the turbine shaft, usually taken in 
revolutions per minute. 

3. Measurement of power with a Prony or a water brake, if 
the power at the turbine shaft is desired ; or with electrical instru- 
ments (ammeters, voltmeters, and wattmeters), if the power is 
measured by the output of an electric generator. 

* For complete and detailed information regarding the testing of steam tur- 
bines and other prime movers, as well as the revised codes of testing adopted by 
the American Society of Mechanical Engineers, "see Power Plant Testing, pages 
294-363, by the author (McGraw-Hill Book Co., N.Y.). 

353 



354 



THE STEAM TURBINE 



4. Weight, or measurement by volume, of the condensed 
steam discharged from the condenser. Unless a surface con- 
denser is used it is very difficult to obtain the amount of steam 
used by the turbine. All leakages from pipes, pumps, and valves, 
which is part of the steam which has gone through the turbine, 
must be added to the weight of the condensed steam. The 
accuracy of a test often depends a great deal on how accurately 
leaks have been provided against, or measured when they occur. 

5. Temperature of the steam as it enters the turbine. If the 
temperature is higher than that due to the pressure of the sat- 
urated steam given in steam tables, the steam is superheated; 
if, however, the temperature is not higher the steam may be 
wet, and a calorimeter must be attached as near the turbine 
steam chest as possible.* 

6. Vacuum or back-pressure. 

All gauges, electrical instruments, and thermometers should be 
carefully calibrated before and after each test so that observations 
can be corrected for any errors. The zero readings of Prony 
and water brakes for measuring power should be carefully 
observed and corrected to eliminate the friction of the apparatus 
with no load. Unless all these precautions are taken the dif- 
ficulties in getting reliable tests of turbines are greatly increased. 
In all cases tests should be continued for several hours with 
absolutely constant conditions if the tests are to be of value. 

The most valuable test of a steam turbine is made when varying 
only the load; that is, with pressures, superheat, and speed con- 
stant. When the steam consumption is then plotted against 
fractions of full load, a water-rate curve is obtained. For such 
a curve a series of tests are needed, each for some fraction of full 
load ; and in each separate test the power as well as all the other 
conditions must be held constant. 

* The most satisfactory tests of turbines are made with steam slightly super- 
heated rather than wet. When steam is very wet (more than about 4 per cent, 
moisture for ordinary pressures) the determination of the quality is difficult. There 
is also a danger that steam showing only a few degrees of superheat by the reading 
of the thermometer is actually wet. The high temperature is due in such cases to 
heating from eddies around the thermometer case or in steam pockets near it. 



TESTS OF STEAM TURBINES 355 

Another important test of the performance of steam turbines is 
made by varying both the speed and the power and keeping the 
other conditions constant. The observations of speed and power 
from such a test give a power parabola as illustrated in Fig. 80. 
This curve shows at what speed the turbine gives the greatest 
output. 

For complete tests of a steam turbine the steam consumption 
should be determined at full load (1) with varying initial steam 
pressure; (2) with varying vacuum; and (3) with varying superheat. 

A complete set of tests as outlined will give sufficient data to 
determine all the corrections usually required. 

Commercial Testing. The methods used by the New York 
Edison Company in commercial tests of steam turbine-generator 
units may well be explained briefly. 

During a test the load on the turbine unit is maintained as con- 
stant as possible by "remote control" of the turbine governor by 
the switchboard operator. The maximum variation in load is 
to be held within 4 per cent, above and below the mean. For 
some time previous to the test the turbine is run a little below the 
load required for the test, but at least ten minutes before the 
starting signal is given the test load must be on the machine. 

Three-phase electrical load is measured by the two-wattmeter 
method,* using Weston indicating wattmeters of the standard 
laboratory type. These instruments are calibrated by a well- 
known testing laboratory immediately before and after the test. 
Power factor is maintained substantially at unity and all electrical 
readings are taken at one-minute intervals. 

When the turbine is provided with a surface condenser, the 
steam consumption, or water rate, is determined by weighing in 
a large tank supported on platform scales the condensed steam 
delivered from the condenser hot well. Above the weighing 
tank a reservoir is provided which is large enough to hold the 
condensation accumulating between the weighings which are 
made at intervals of five minutes. By using a loop connection 

* Cf. Kent's Mechanical Engineer's Pocket-Book, 7th ed., page 1069, 8th ed., 
page 1396, or Foster's Electrical Engineer's Pocket-Book. 



356 THE STEAM TURBINE 

for the gland water supply (of Westinghouse turbines) or the 
water from the step bearing (of Curtis turbines using water for 
this bearing) the necessity for correcting the weighings for these 
amounts is avoided. 

Because the circulating water at the stations of this company 
is usually quite salt, any condenser leakage is detected by testing 
the condensed steam by the silver-nitrate method with a suitable 
color indicator. This color method is said to be a decided 
advantage over the usual method of weighing the leakage accu- 
mulating during a definite period when the condenser is idle and 
is tested for only one particular vacuum. By taking samples 
of circulating water and condensed steam at the same time, 
it is possible to detect any change in the rate of condenser 
leakage. 

The water level in the hot well is maintained at practically a 
constant point by means of a float valve in the well automatically 
controlling the speed and, therefore, the amount of the delivery 
of the hot-well pump. This device avoids the necessity for the 
difficult correction to be made in a test when the levels in the hot 
well are not the same at the beginning and end of a test. Tem- 
peratures and pressures of the admission steam are determined 
by mercury thermometers and pressure gauges located near the 
main throttle valve of the turbine; the amount of superheat is 
determined by subtracting from the actual steam temperature 
after making thermometer corrections the temperature of 
saturated steam corresponding to the pressure at the point where 
the temperature is measured. All gauges and thermometers are 
calibrated before and after the test. 

Vacuum is measured directly at the turbine exhaust by means 
of a mercury column with a barometer alongside for reducing the 
vacuum to standard barometer conditions (30 inches). By this 
latter arrangement the necessity for temperature corrections 
which are necessary when the two mercury columns are not at 
the same place is avoided. 

Fig. 188 shows a 5500-kilowatt Westinghouse-Parsons turbine 
set up for testing in the shops before shipment to the customer. 



TESTS OF STEAM TURBINES 



357 




358 



THE STEAM TURBINE 



The power is measured by means of a large water brake shown in 
the figure at the left of the turbine. 

Reports of Tests. The tables given below have been prepared 
to show the steam consumption, together with the most impor- 
tant other data, of what are believed to be reliable tests of 
standard makes of steam turbines. The vacuum given in the 
tables is the equivalent referred to 30 inches barometer. 

Curtis Turbines. The following results were obtained in 
1905 by Messrs. Sargent and Lundy with a 2000-kilowatt Curtis 
turbine-generator. 



Kilowatts. 


Steam Pressure 
(Gauge). 


Superheat, 
Deg. F. 


Vacuum, Inches. 


Pounds per 
Kilowatt-hour. 


555 
1067 
2024 


155-5 
170.2 
166.3 


204 
120 
207 


28.5 
28.4 

28-5 


18.09 
16.31 
15.02 



Also the following results are reported in 1907 with a 9000- 
kilowatt turbine-generator in Chicago : 



Kilowatts. 


Steam Pressure 
(Gauge). 


Superheat, 
Deg. F. 


Vacuum, 
Inches. 


Pounds per 
Kilowatt-hour. 


5,374 

8,070 

10,186 

13,900 


182 
179 
176 
198 


133 
116 

147 
140 


29-43 
29-35 

29.47 
29.31 


I3- J 5 
13.00 
12.90 
13.60 



Parsons Turbines. A 1500-kilowatt Parsons turbine was 
tested at Sheffield, England, with the following results: 



Kilowatts. 


Steam Pressure 
(Gauge). 


Superheat, 
Deg. F. 


Vacuum, 
Inches. 


Pounds per 
Kilowatt-hour. 


530 
1071 

1585 


145 -° 
131. 

128.5 


no 
124 
125 


28.9 
28.3 

27-5 


21.58 
18.24 
17.60 



TESTS OF STEAM TURBINES 



359 



The results of two tests of a 300-kilowatt Parsons turbine 
installed at the Hulton colliery are also given to show the change 
of economy from running condensing 26.58 inches vacuum and 



non-condensing. 



Kilowatts. 


Steam Pressure 
(Gauge). 


Superheat, 
Deg. F. 


Vacuum, 
Inches. 


Pounds per 
Kilowatt-hour. 


3°3 
297 


158.0 
161. 






26.6 
0. 


23-I5 
34-20 



These last tests show well the increased steam consumption 
(about 50 per cent. ) when running non-condensing. 

Westinghouse-Parsons Turbines. The table below gives the 
results of tests in 1904 by F. P. Sheldon & Co., Providence, 
R.I., of a 400-kilowatt Westinghouse-Parsons turbine with about 
100 degrees F. superheat. 



Brake 
Horsepower. 


Steam Pressure 
(Gauge) . 


Superheat, 
Deg. F. 


Vacuum, 
Inches. 


Pounds per 
B.H.P. Hour* 


279.4 
410.7 

657-3 

967-5 
1207.5 


I53-I 
153-2 
I52-7 
149.6 
152.0 


92-5 
102.9 
100.3 
100.2 

99.9 


28.0 
28.0 
28.0 
27.6 
27;3 


14-34 
13-45 
12.48 
12.79 
13-55 



* Observe the steam consumption is in pounds per brake horsepower hour, instead of pounds per 
kilowatl-hoxa as for some of the other results given here. 

The curves given in Fig. 189 were plotted to show graphi- 
cally the steam consumption of 300, 500, and 1000 kilowatt 
Westinghouse-Parsons turbines with varying loads.* Data of 
the tests from which these curves were drawn, as well as of a 
test of a 3000-kilowatt turbine are given in the following tables. 
These tests were reported by J. R. Bibbins in 1906 and 1907. 

* The numbers marked on the curves to indicate the vacuum represent the 
actual readings taken in the test and are not referred to a standard (30 inches) 
barometer. 



360 



THE STEAM TURBINE 




o .5 
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ii 

a. ° 
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c o 

o £ 



1 u 



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300-KILOWATT TURBINE (3600 R.P.M.). 



Brake 
Horsepower. 


Steam Pressure 
(Gauge) . 


Superheat, 
Deg. F. 


Vacuum, 
Inches. 


Pounds per 
B.H.P. Hour. 


232.7 
460.6 
688.5 


145- 1 
144.6 
140.3 


4.1 

4.8. 

7.0 


28.0 
28.0 
27.2 


15-99 
13-99 
15-73 



TESTS OF STEAM TURBINES 361 

500-KILOWATT TURBINE (CONDENSING AND NON-CONDENSING) (3600 R.P.M.). 



Brake 
Horsepower. 


Steam Pressure 
(Gauge). 


Superheat, 
Deg. F. 


Vacuum, 
Inches. 


Pounds per 
B.H.P. Hour. 


383-5 

755-6 
1121.9 


152.6 
149-2 
148.8 


.2 

1.2 


28.2 
27.8 
26.5 


14-15 
13.28 

'I4-3 2 


385.6 
766.8 

1144.4 


148.2 

147-3 
126. 1 


2.7 

2.6 
11. 4 


0.8 
0.8 
0.8 


24.94 
22.IO 
24.36 



These last tests show well the increased steam consumption 
(about 75 per cent.) when running non-condensing. 



1000-KILOWATT TURBINE (1800 R.P.M.). 






Brake 
Horsepower. 


Steam Pressure 
(Gauge). 


Superheat, 
Deg. F. 


Vacuum, 
Inches. 


Pounds per 
B.H.P. Hour. 


752.4 

I503-5 
2252.7 


i5°-5 

146.7 
145-3 


0.2 

O.O 
0.0 


27-5 
27.O 
25.2 


14-77 
13.61 

I5-29 




3000-KIL 


OWATT TURBINE 


(1500 R.P.M.). 




Brake 
Horsepower. 


Steam Pressure 
(Gauge). 


Superheat, 
Deg. F. 


Vacuum, 
Inches. 


Pounds per 
B.H.P. Hour. 


2295 
4410 


152.0 
143-9 


102 

87 


26.2 
26.2 


12.36 
II.85 



Rateau Turbine. A iooo-kilowatt Rateau turbine built at the 
Oerlikon works gave the following results of steam consumption : 



Kilowatts. 


Steam Pressure 
(Absolute). 


Superheat, 
Deg. F. 


Vacuum, 
Inches. 


Pounds per 
Kilowatt-hour. 


194 

425 

871 

1024 


186 

155 
181 
179 


47 

21 
11 
10 


27-73 

27.6 

23.6 

25 -°5 


31-97 
24.91 
24.69 
21.98 



362 



THE STEAM TURBINE 



Zoelly Turbine. A 5500-kilowatt Zoelly turbine installed at 
the Ouest Electricity Works, Paris, is said to operate at full load 
with a steam consumption of approximately 12.0 pounds per 
brake horsepower-hour at 160 pounds per square inch gauge 
pressure, 200 degrees F. superheat, and 27 inches vacuum. 

De Laval Turbine. The following table gives results of tests 
by Dean & Main of a 300-horsepower De Laval turbine: 



Brake 
Horsepower. 


Steam Pressure 
(Gauge). 


Superheat, 
Deg. F. 


Vacuum, 
Inches. 


Pounds per 
B.H.P. Hour. 


196.O 
298.4 

352- 


197.7 
197.O 

198. 5 


16 
64 
84 


27.4 
27.4 
27.2 


15.62 
14-35 
13-94 



The results shown in the above tests give the relative steam 
economy of the principal types of turbines from light load to 
overload. Tables I and II * on the following page give the com- 
parative results of the latest reported tests in America and in 
Europe. 

HEAT UNIT BASIS OF EFFICIENCY. 

The usual methods used for correcting steam turbine tests to 
get. a standard for comparison explained in Chapter VI are not 
established on a highly scientific basis. Engineers appreciate 
generally that a more rational method of comparison of the 
economy of heat engines on a heat unit basis should be adopted 
in cases where it is practicable. As regards steam turbines there 
are, however, so many uncertain factors entering into the deter- 
mination of a thermodynamic efficiency from the available energy 
that for the present such methods can be of little value, except 
in some special cases. Comparatively high superheats are now 
generally used, and our knowledge of the effect of reheating in a 
multi-stage turbine is very indefinite. 

A thermal efficiency can, however, be calculated readily and 

more satisfactorily by determining what percentage the heat 

equivalent of the work is of the heat "used by the turbine," 

assumed to be the difference between the total heat in the steam 

* Compiled by H. T. Herr and A. G. Christie. 



TESTS OF STEAM TURBINES 



363 





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364 THE STEAM TURBINE 

at the initial conditions and the heat ("of the liquid") in the 
condensed steam at the temperature of the exhaust. 

By this method the full load test of a Westinghouse-Parsons 
turbine reported by F. P. Sheldon & Co. will be calculated from 
the data given in an official report. 

In order to make the results of such calculations of steam 
turbine tests comparable with the usual heat unit computations 
of reciprocating steam engine tests the results are generally 
expressed in terms of indicated or " internal " horsepower. 
F. P. Sheldon & Co. assumed the mechanical efficiency of a 
reciprocating engine of about the same capacity at full load to 
be 93.3 per cent. 

THERMAL EFFICIENCY OF A 400-KILOWATT STEAM TURBINE. 

Brake horsepower 660 

Corresponding indicated or "internal" horsepower of a recip- 

660 
rocating engine = 708 

Total steam used per hour, pounds 9169 

Steam used per " internal" horsepower per hour, pounds .... 12.96 

Steam pressure, pounds per square inch absolute 166.9 

Superheat, degrees F 2.9 

Vacuum, referred to 30 inches barometer, inches 28.04 

Temperature of condensed steam, degrees F. (at .96 pound per 

square inch absolute pressure) 100. 6 

Total heat contents of one pound of dry saturated steam at 

the initial pressure, B.T.U 1193 . 9 

Heat equivalent of superheat in one pound of steam, B.T.U. 

{Cp. from Fig. 30) 1.9 

Total heat contents of one pound of superheated steam, B.T.U. 1195.8 

Heat of liquid in condensed steam, B.T.U 68.6 

Heat used in turbine per pound steam, B.T.U n 27. 2 

Heat used in turbine per "internal" horsepower per 

minute, B.T.U. = 1127.2 X —j$- = 243.5 

Heat equivalent of one horsepower per minute, B.T.U. = _ 42 .42 

778 

Thermal efficiency, per cent. (42.42 -j- 243.5) J 7-4 

Standard forms for data sheets and for tabulating results of 
steam turbine tests are given in Power Plant Testing by the 
author. (See pages 315-340.) Full explanations of methods 
and of necessary precautions are given. 



CHAPTER XIV. 
STEAM TURBINE ECONOMICS. 

The Best Conditions of Vacuum, Superheat, and Steam Pressure. 

For normal operating conditions, a great deal can be learned 
about the most profitable and satisfactory vacuum, superheat, 
and initial steam pressure for steam turbines from a comparison 
and study of existing modern power plants. 

For this purpose a table * is given on the following page in 
which data are given regarding the vacuum, superheat, and steam 
pressure of a large number of steam power plants. This table 
is compiled from fifty-eight turbine plants in America and in 
England. The figures represent the number of plants working 
under the conditions stated at the head of each column. 

There is no doubt that such comparative data of operating 
conditions are, from a practical viewpoint, of considerable impor- 
tance. Although these figures were collected in 1904 and 1905, 
they may be taken to represent very well the average practice of 
the last few years as well, except that in America there has been 
a tendency to operate a larger percentage of the plants with 
from 100 to 150 degrees F. superheat and at about 28 inches 
vacuum. 

The Question of the Most Profitable Vacuum. Steam turbine 
manufacturers are inclined, naturally, because of the obvious 
advantages of turbines over reciprocating engines for operation 
at high vacuums, to draw attention to the reduction in the steam 
consumption when a plant is operated at a high vacuum. Then 
the question is often raised as to the actual economy considering 
the increased first cost of the condensers, pumps, and piping, 

* J. R. Bibbins, in the Report of American Street Railway Association, October, 
1904, page 201, and from data collected in 1905 by Messrs. Stevens and Hobart. 

36S 



3 66 



THE STEAM TURBINE 



together with probably larger operating expenses. A great deal 
depends on the local conditions, particularly on the average 
temperature of the condenser cooling water. At places only 
slightly elevated above the sea-level and where the temperature 
of the water supply for the condensers is very low — near the 



Limits of 
Capacity in 


Character of 
Service. 


Limits of 
Vacuum, 
Inches of 
Mercury. 


Limits of 
Superheat, 
Degrees F. 


Limits of 

Steam Press., 

Pounds Gauge. 


Rated Kilo- 
watts of Plant. 


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10,000 to 5,000. 
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[Traction.. . . 
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Traction.. . . 

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freezing point for a large part of the year — it is doubtless profit- 
able to install condensing apparatus of sufficient size to operate 
steam turbines at from 28.5 to 29 inches vacuum. The following 
table, calculated by J. R. Bibbins, gives side by side the 
theoretical and the practical vacuums at sea-level for varying 
temperatures of the cooling water. 



STEAM TURBINE ECONOMICS 



367 



VACUUM AT SEA LEVEL FOR VARYING TEMPERATURES OF COOLING WATER. 



Temperatures of Cooling 
Water. Deg. F. 


Theoretical 
Possible 
Vacuum. 
Inches. 


Perfect Con- 
denser, No 

Temperature 

Difference. 

Inches. 


Actual Con- 
denser, 15 F. 
Difference. 
Inches. 


Actual Con- 
denser, 15 F. 
Difference. 
Inches. 


Ratio Water to Steam 


Infinite. 


60 to 1. 


60 to 1. 


100 to 1. 


32 
60 
70 
75 


29-83 
29.50 
29-30 
29.IO 


29.67 
29.12 
28.77 
28.51 


29-43 

28.56 

27. 72 
27-37 


29-54 
28.82 
28.38 
28.11 



In modern surface condenser installations there is usually a 
difference of about 15 degrees F .* between the temperature of the 
condensed steam and of the discharged water. It will be seen 
then in the above table that with the reasonable ratio of cooling 
water to steam of 60 to 1 the maximum vacuum obtainable, when 
the cooling water is taken in at 60 degrees, is 28.6 inches, and 
when taken in at 70 degrees is only 27.7 inches. f 

The fact must not be lost sight of that the elevation has an 
appreciable effect on the maximum possible vacuum and con- 
sequently on the most profitable vacuum. At an elevation of 
1000 feet above the sea-level the possible vacuum obtainable 
with a given condensing apparatus will be about an inch less than 
at tide-water, and the vacuum reduction is, of course, in pro- 
portion for other elevations. 

Bibbins has also calculated the actual percentage saving when 
the condenser equipment is increased so that the plant can be 
operated at 28 inches instead of 26 inches. It is estimated that 
the cost of the condenser equipment including pumps and piping 
will be $4000 more for a 2000-kilowatt plant to operate at 28 
inches vacuum than at 26 inches vacuum. The results are given 
in the table on the following page: 

* This is a conservative value. In good modern practice a considerably smaller 
temperature difference is obtainable. 

t A firm of engineers, which has been installing steam turbines almost exclusively 
in the power plants it has designed and constructed, has equipped a power plant 
at Tampa, Florida, with Diesel oil engines because of the cost of cooling water in a 
warm climate. 



3 68 



THE STEAM TURBINE 



RELATIVE ECONOMY OF 28 INCHES VACUUM OVER 26 INCHES IN A 
200O-KILOWATT PLANT. 

Estimated Increased Cost of Equipment is $4,000. 



Net Saving 
expressed as 
Percentage of 
Increased Cap- 
ital Cost to 
Secure 28 Ins. 
Vacuum over 
that for 26 Ins. 


Average 

Load 

in 

Kilowatts. 


Hours of 

Service 

per 

Day. 


Actual 
Evapo- 
ration, 
Pounds. 


Steam 
Consump- 
tion, 
Average 
Pounds 
per Kilo- 
watt-hour. 


Water 
Saved per 
Kilowatt- 
hour 
by Rais- 
ing Vac- 
uum from 

26 Ins. 
to 28 Ins. 


Coal, 

Dollars 

per 

Ton. 


118 

27 
4 


1500 
1000 
1000 


24 
24 
10 


9-5 

8 

8 


23 
22 

22 


I.84 
1.76 
I.76 


4-5° 

2.25 
I-I3 



In the calculations for the above results the rate of interest was 
taken at 5 per cent, and depreciation at 7.5 per cent, on the extra 
cost of equipment. Cost of extra power consumed was at the 
rate of 1 cent per kilowatt-hour, and 10 cents per 1000 gallons of 
feed-water saved. 

Although it may be stated, in general, that it is profitable to 
equip a station to operate under normal conditions at a vacuum 
of 28 inches instead of 26 inches, it will be observed from the 
above table that there are cases where there is practically no 
advantage either way. In the third case given, where the plant 
has only a 10-hour load and coal is cheap, the gain is only 4 per 
cent. 

Operation at 29 inches vacuum compared with 28 inches is 
not nearly so favorable to the higher vacuum as the comparison 
of 28 inches with 26 inches. 

It will be observed in the table on the following page that the 
volume of the steam is increased practically in the same ratio (the 
volume is practically doubled) when the vacuum is increased from 
28 inches to 29 inches as when increased from 26 inches to 28 
inches. Fig. 181 shows graphically the very large increase in 
volume of the steam in its passage through the five stages of a 
large Curtis turbine operating at 29 inches vacuum. 



STEAM TURBINE ECONOMICS 



369 



TABLE OF THE VOLUME OF A POUND (SPECIFIC VOLUME) 
OF DRY SATURATED STEAM AT HIGH VACUUMS. 



Vacuum, Inches. 


Volume, Cubic Feet. 


29 
28 
26 


665* 

342 

176 



* Ratio of the volume at 28 inches vacuum to that at 26 inches is 1 .94, and the ratio of volumes 
at 29 inches vacuum and at 28 inches is 1 .95 . 

It may be stated then that the capacity of the condensing 
equipment for a turbine operating at 29 inches vacuum must be 
practically four times as large as it would be for one exhausting 
at about 26 inches vacuum. Or, in other words, the volume of 
a pound of steam at the exhaust is 166 cubic feet larger at 28 
inches vacuum than at 26 inches, but that at 29 inches it is 323 
cubic feet larger than at 28 inches. Now if as has been stated 
the cost of the condensing equipment is $4000 more for a 2000- 
kilowatt unit when 28 inches vacuum is substituted for 26 inches, 
the increased cost is obviously much greater when 29 inches 
vacuum is compared with 28 inches. 

For turbines of which the steam consumption is not reduced 
very much more per inch of vacuum between 28 and 29 inches 
than between 26 and 28 inches, in a comparison of the economic 
operation at 29 inches vacuum with 28 inches there is a large 
increased capital cost for condensing equipment which is not 
offset by a proportionate reduction of the steam consumption, 
and there are probably comparatively few places — unusually 
located as regards low elevation, low temperatures, large capacity, 
expensive fuel, or high load factor — where an installation for 
operation at an average vacuum of 29 inches is profitable. 

The percentage change in the steam consumption is approxi- 
mately the same at light ("fractional'') loads as at full load (see 
page 166). Now because the steam consumption per kilowatt- 
hour is greater at light loads, the change in steam consumption 
per kilowatt-hour is therefore also greater at light loads than at 



37° 



THE STEAM TURBINE 



full load. Ordinarily this fact is stated by saying that a change 
in vacuum has a greater "effect" at light than at full load, and 
that the effect is more marked at high than at low vacuums. 
The effect of vacuum on the steam consumption of any impulse 





























































J 


































































4.0 


























































/ 


























































1 






a 3.5 

s 
























































/ 


























































/ 








£ 3.0 






















































/ 


























































/ 










I 2 ' 5 




















































/ 






































































2 
a 2.0 

H 




























































































, 






























1„ 


























































































































* 1.0 


























































































































.5 


























































































































a 
























































\ 





2 4 6 8 10 12 14 16 18 20 22 24 26 2S 
Vacuum, in Inches of Mercury 

Fig. 190. Percentage Curve of the Effect of Vacuum on the Steam Consumption 
of a Single-Stage Impulse Turbine. 



turbines of the single-stage type is probably shown very accurately 
by Fig. 190 (reproduced from Fig. 88).* 

* In the catalogs of the General Electric Company it is stated that a curve like 
Fig. 127 is typical for most Curtis turbines. Actually, however, a curve like 
Fig. 78 is more accurate. Emmett has stated recently in a published communi- 
cation that "around 27 inches the change in economy per inch is 6.6 per cent.; 
28 inches 7.8 per cent.; and 29 inches 9.5 per cent." 

Parsons states in a paper read before the Institution of Electrical Engineers in 
1904 that in a turbine "the benefit derived from a good vacuum is much more than 
in a reciprocating engine. Every inch of vacuum between 23 and 28 inches affects 
the steam consumption on an average about 3 per cent, in a 100-kilowatt; 4 per cent, 
in a 500-kilowatt; and 5 per cent, in a 1500-kilowatt turbine, the effect being more at 
high vacuum and less at low." It seems very doubtful to the author whether, in 
general, vacuum corrections can be classified according to the size of the turbine. 
There are some very large turbines of the Parsons type of which the vacuum 
correction is less than 4 per cent, per inch of vacuum. 



STEAM TURBINE ECONOMICS 



37 1 



The variation of the steam consumption of a 500-kilowatt 
Westinghouse-Parsons turbine for vacuums from 25 to 29 inches 




s 



i£ 



iU 



Fig. 191. Curves of Steam Consumption of a 500-Kilowatt Westinghouse-Parsons 
Turbine for 25, 26, 27, 28, and 29 Inches Vacuum. 



from light loads to overloads is illustrated by the curves in Fig. 191. 
What might be called a curve of normal vacuum correction 



372 



THE STEAM TURBINE 



factors for comparing those of 26, 27 and 29 inches with 28 inches 
in Westinghouse turbines is given in Fig. 192. 

Chilton,* after stating that the impression is no longer so 
common that a high vacuum is necessary to secure good results 
with steam turbines, says that the difference in economy of Allis- 
Chalmers-Parsons turbines between 24 and 27 inches vacuum 
is 5 per cent, per inch. Between 27 and 28 inches the saving is 
6 per cent., and between 28 and 29 inches is 7 per cent. 






Fig. 192. Vacuum Correction Factors for Westinghouse Single-Flow Turbines. 



An idea of the relative quantity of condensing water required 
for different vacuums may be gained by comparing that required 
for the usual operating vacuums. For example, with injection 
water of 70 degrees F., the usual temperature upon which con- 
denser guarantees are based, it is customary to estimate that 
to obtain a vacuum of 27 inches about 36 pounds of water will be 
used for each pound of steam condensed, and about 1.4 times this 
quantity is required for a vacuum of 28 inches. With injection 
water at 60 degrees F., which may be considered the winter tem- 
perature, the quantities required for the foregoing vacuums are 

* Street Railway Journal, Oct. 19, 1907. 



STEAM TURBINE ECONOMICS 



373 



approximately 28 and 34 pounds respectively. Having the quan- 
tity of condensing water required, the cost of fuel, and cost of 
water delivered to the condenser, the vacuum best suited to the 
conditions under consideration may be readily determined. Theo- 
retically, the effect upon the turbine of reducing the vacuum 
below that for which it is designed, is to reduce the capacity and 
to lower the rating at which maximum economy is obtained. 

The following table * illustrates the percentage gain in economy 
per inch of vacuum for various vacuums. The close agreement 
between the actual results and the theoretical values should be 
observed. The table applies, however, only to turbines using 
very high steam pressures and superheats. For " land " turbine 
practice it is serviceable only for comparison. 



Inches of Vacuum. 


Gain in Per Cent. 


28 


27 


26 


25 




5-i 
5-° 
3-14 
5-2 


4-8 
4.0 

3-°5 
4.4 


4-6 

3-5 
2.95 

3-7 


4.2 

3-° 

2.87 
30 






Theoretical 





Effect of Superheating on Economy. The effect of superheat 
on the economy of De Laval and Parsons turbines is usually 
stated to be 10 per cent, per 100 degrees F. superheat. This 
statement is probably very nearly correct for the usual ranges of 
superheat in practice and is the usual correction employed by 
most consulting engineers for correcting steam turbine tests f up 
to about 150 degrees F. superheat. 

Some investigations made by Professor Hobart show that the 
mean superheat correction for Parsons turbines is almost exactly 

* Mechanical Engineer, Feb. 24, 1906. 

f The superheat corrections used by the engineers of the Westinghouse Com- 
panies, by Dean & Main, and by Parsons, are all approximately 10 per cent, per 100 
degrees F. superheat. 



374 THE STEAM TURBINE 

10 per cent, per ioo degrees superheat for all superheats from 
o to ioo degrees F. Between ioo and 150 degrees superheat it 
is approximately 8 per cent., and between 150 and 250 degrees 
is about 6 per cent. It is the opinion of the author that the 
results of this investigation can be considered quite accurate, as 
a large number of tests were compared. A curve showing 
approximately the same sort of variation in the superheat correc- 
tion of De Laval turbines is given in Fig. 87. Chilton states that 
tests of Allis-Chalmers-Parsons turbines show that the " incre- 
ment of saving becomes smaller as the superheat is increased"; 
adding that for 50 degrees F. superheat the steam consumption 
is reduced 7 per cent, (at the rate of 14 per cent, per 100 degrees) ; 
for 100 degrees 10 per cent; and for 150 degrees 12.5 per cent, 
(at the rate of a little more than 8 per cent, per 100 degrees). 

According to Kruesi of the General Electric Company 100 
degrees F. superheat reduces the steam consumption of Curtis 
turbines 8 per cent., but "the first 50 degrees of superheat is of 
greater value than the second 50 degrees."* 

When steam at about 150 pounds per square inch gauge 
pressure is superheated 100 degrees F. the total heat of the 
steam is increased about 4.8 per cent, with an additional fuel 
expenditure of approximately 6 per cent, if the boiler equipment 
is good. Now since the steam consumption is reduced from 8 to 
10 per cent, for 100 degrees F. superheat there is obviously a 
saving of from 2 to 4 per cent, in the cost of fuel. 

Experience seems to show that the best economic results will 
be obtained with from 100 to 150 degrees F. superheat for tur- 
bines of the Parsons type, and about 50 degrees superheat for 
Curtis turbines of more than one stage. In all kinds of turbines 
of the single-stage impulse type there is probably always a saving 

* Because curves of steam consumption per kilowatt-hour for varying super- 
heats (like Fig. 126) were apparently straight lines, most turbine engineers, until 
very recently, believed that at high superheats the percentage correction was 
increased instead of being reduced as more recent results show. Since it has 
been fairly well established that the specific heat of superheated steam has very low 
and minimum values at from about 200 to 250 degrees F. superheat, the later 
results seem to be the more reasonable. 



STEAM TURBINE ECONOMICS 375 

of from 4 to 5 per cent, in fuel cost per 100 degrees F. superheat 
within the practicable limits of superheating. 

Although there is much yet to be determined concerning 
superheated steam, it has been shown by experience in turbine 
plants that a considerable saving in fuel can be secured by super- 
heating the steam at least a moderate amount. The greater 
saving in turbines of the Parsons type over multi-stage Curtis 
turbines is due to the larger " skin-friction " or disk and blade 
rotation losses of the large number of rows of blades in Parsons 
turbines. The curves in Fig. 69 show the very large percentage 
that these losses are reduced when the blades revolve in dry 
steam instead of wet steam. When the admission steam to a 
Parsons turbine is dry saturated the steam in the low-pressure 
stages will probably have, nearly 20 per cent, of moisture, while 
if it is superheated 150 to 200 degrees F. the steam in these stages 
will be nearly dry. 

Finally, the use of a high degree of superheat must depend not 
only on the type of turbine, the load factor, and the size of the 
units but also upon the nature of the service as regards severe 
and frequent variations in the load, having in mind the difficulties 
which have been encountered in the practical operation of super- 
heaters, steam piping, valves, pumps, and auxiliary machinery. 

Reasons for the Improved Economy in Turbines and Recipro- 
cating Engines Due to Superheated Steam. A gain in steam and 
fuel economy results from the use of superheated steam in either 
turbines or reciprocating engines. In the turbine the gain comes 
principally from the reduced fluid friction of the steam moving 
at a high velocity through passages and blades, some of which 
have also a comparatively high velocity. In a reciprocating 
engine the gain from superheated steam is due to the reduction of 
cylinder condensation, resulting in less loss due to the cooling of 
the cylinder from the reevaporation of moisture at the lower 
pressures near the end of the stroke. On account of this cooling 
of the cylinder ends, the loss due to the "initial condensation" 
of the steam admitted on the return stroke is often 40 to 50 per 
cent, of the weight of steam admitted. This loss is partly or 



376 THE STEAM TURBINE 

entirely prevented when the steam is superheated, depending 
upon the degree of superheat. In a steam turbine there is a 
similar loss due to condensation, but it is due almost entirely to 
the mere expansion of the steam. The walls of the turbine 
casing remain, however, at a practically uniform temperature, so 
that there is no opportunity for loss through reevaporation of 
condensed steam. 

Steam Pressure Best Suited to Turbines. It is the general 
opinion of practical engineers that probably the most economical 
operating pressure for the usual power-house services is about 
150 pounds per square inch gauge pressure (165 absolute) at the 
throttle valve, and that a greater saving can always be obtained 
by the use of a moderate amount of superheat than by increasing 
the pressure beyond this point. 

Chilton states that there is a gain of 2 per cent, in steam con- 
sumption from increasing the steam pressure from 150 to 175 
pounds per square inch* and 1 per cent, for an increase from 
175 to 200 pounds per square inch. But against the saving in 
fuel due to a reduced steam consumption must be charged the 
increased cost of piping, valves, and boilers, and also the loss due 
to increased leakage. Increasing the steam pressure will also 
increase considerably the cost of the turbine. A "rough and 
ready" correction used a great deal by turbine engineers is one- 
tenth per cent, per pound. 

Speed Variation as it Affects Economy. A steam turbine will 
give its best economy at some particular speed,f just as it has been 
found to give its best economy at some definite load. For this 
reason the design of a turbine should be worked out very carefully 
with velocity diagrams to determine whether at the speed required 
by the operating conditions it will give the best economy. When- 
ever any changes are made in the design of a turbine, the 
manufacturers will always make tests to determine the steam con- 

* The engineers of the Westinghouse and General Electric companies use prac- 
tically the same correction for initial pressure. It may be added that the correc- 
tion for exhaust pressure (back pressure) of non-condensing turbines is about ten 
times as large as the correction for initial pressure. 

f See curves on page 174. 



STEAM TURBINE ECONOMICS 377 

sumption at various speeds, and curves like those shown in Fig. 80 
are calculated and plotted. If it is found that the turbine has a 
lower steam consumption at a slightly different speed from that 
for which it is rated, either the angles of the blades or the pressures 
must be changed. The reasons for such changes are obvious, 
because the blade speed has a very definite relation to velocity 
of the steam in the blades. If the designer is not successful in 
securing this relation for the rated speed, there will be impact of 
the steam against the blades and a consequent loss of efficiency. 

The curve of steam consumption in Fig. 80 shows the change 
in economy at various speeds. At 2000 revolutions per minute 
the steam consumption is 19.6 pounds per kilowatt-hour; at 
1800 revolutions (rated speed) it is 19.45 pounds; at 1600 
revolutions, about 19.8 pounds; at 1400 revolutions, about 20.7 
pounds; and at 1000 revolutions, about 24.7 pounds. It will be 
observed in these curves that the ideal conditions have been 
secured in the design of this turbine; that is, the steam consump- 
tion is lowest and the output (load) greatest at the rated speed. 
Within a range of about 50 revolutions above or below the rating 
(a total variation of about 6 per cent.) the steam consumption is 
practically constant. These curves are typical for all good 
designs of steam turbines. 

When a speed test is made of an impulse turbine the best 
results are obtained as regards the accuracy of the design by 
running the turbine with a number of nozzles wide open to give 
approximately full load. The test for each speed can then be 
made of comparatively short duration, as the steam can be weighed 
continuously between the first and last tests without interruption 
when the speed is being changed. With a constant number of 
nozzles discharging steam the rate of flow will be the same at all 
speeds. 

Comparative Economy of Steam Turbines and Reciprocating 
Engines. To summarize the results of tests on a number of 
large steam turbines and reciprocating engines the following 
tables have been prepared. Steam consumption of most of the 
turbine tests was given in the published data in terms of kilowatt- 



378 



THE STEAM TURBINE 



hours or electrical horsepower-hours.. In order to make com- 
parisons with the reciprocating engine it was necessary to reduce 
all to a common standard — brake horsepower-hour. To express 
all the results in this common standard various efficiencies must 
be assumed. In the calculations the generator efficiencies given 
on page 454 were used to obtain the following coefficients to 
change the steam consumptions from the rate per kilowatt-hour 
to that per brake horsepower-hour: 



Rating of Turbine, 
Kilowatts. 


Coefficient. 


300 and 400 

500 
1,000 to 3,000 
5,000 to 10,000 


.68 

•71 
.72 

■73 



Mechanical efficiency of reciprocating engines of 3000 to 5000 
horsepower is about 91 per cent.; 1000 horsepower, about 90 per 
cent. ; and 400 to 700 horsepower, about 89 per cent. 

In the following tables are given the steam consumptions of a 
large number of steam turbines and some particularly good 
reciprocating engines. A great many of the steam turbine tests 
given are approximately the full load data taken from the tests 
recorded at the end of the preceding chapter, and some others are 
taken from Chapter VI. 

The ratings given in the tables are those for what is generally 
known by engineers as "full load;" meaning that the turbine 
can carry economically a load at least 50 per cent, larger than 
this rating. This statement is necessary because some manu- 
facturers use a rating based on maximum output. 

Assuming average values of the corrections given above by 
various authorities, an approximate equivalent steam consumption 
has been calculated for each engine at o degrees F. superheat, 
28 inches vacuum, and 165 pounds per square inch absolute 
steam pressure. 



STEAM TURBINE ECONOMICS 



379 



STEAM CONSUMPTION OF TURBINES. 



A «= American, E = English, F 



French, G = German, S = Swiss, and W-P = Westinghouse- 
Parsons. 





Rated 
Power. 


Conditions of Test. 


Steam per Horn- 
as per Test. 


Equivalent 
Steam per 
b.hp.-hr. at 


Turbine. 


Super- 
heat, 
Deg.F. 


Vacuum 
Inches. 


Steam 

Pressure 

Lbs. 

Abs. 


r.p.m. 


Pounds 
per kw. 


Pounds 
per 
b.hp. 


Degs. 
Sup., 28 ins. 
Vac, 165 
Lbs. Abs. 

Press.* 


De Laval (G).... 


hp. 

30 

150 

300 

kw. 
300 
300 
300 

hp. 
500 

kw. 

500 

500 

500 

500 

1000 

1000 

1200 

1500 

2000 

3000 

5000 

7500 

9000 








5 
100 

107 

1 

290 

104 





10 

468 

125 

207 

235 

142 

96 

116 


non-con. 
26. 4 
26.6 

26.6 
28.0 
28.0 

28.7 

27.8 
28.0 
26. 9 
26. 7 
27.0 
25. 
28.8 
27.5 

28. s 
27.0 
28.8 
27.3 

29. 6 


100 
114 
206 

158 
160 
168 

201 

164 
165 
168 
136 
163 
179 
178 
144 
181 
i39 
189 
193 
194 






39.6 

17.70 

15.17 

15.70 
13-99 
12.48 

13.37 

13.28 

10. 71 
14.55 
15.05 
13.61 
15.80 

11. 00 

12. 67 
10.82 
10. 60 

9.87 




De Laval (G).... 






15.81 
15.29- 

14. 71 
13.99 
13.77 

15.84 

13.17 
13. 20 
14.68 
13-25 
13.04 
12.83 
15.52 
13.48 
13.47 
12. 00 
n.95 


De Laval (A).... 






Parsons (E) 

W-P (A) 


3000 
3600 
3600 

3606 
1800 
1800 
2400 
1800 

900 
1350 
75o 
7 So 
75o 


23.15 

18.82 

15.10 
20.5 
21. 2 

21.98 
15.30 
17. 60 
15.02 
14.74 
13-52 
15.15 
13-00 


W-P (A) 


Zoelly(G) 

W-P (A) 


Curtis (A) 

Curtis (E) 

Rateau (F) 

W-P (A) 


Rateau (S) 

Parsons (S) 

Parsons (E) 

Curtis (A) 

Parsons (G) 

Curtis (A) 

W-P (A) 


Curtis (A) 


9. 11 12.00 



Additional data from recent tests of steam turbines are given 
in tables I and II, page 363. 

Combined Steam Engine and Low-pressure Steam Turbine. 
The combination units of a large Allis engine with Curtis ex- 
haust steam turbines (see Fig. 181c, page 317), as installed in the 
59th Street Power Station of the Interborough-Metropolitan 
System in New York, have a rated capacity of 15,000 horsepower 

* Correction curves in Figs. 87 and 88 were used to correct the De Laval tests for superheat and 
vacuum and the usual correction of .1 per cent, improvement in economy per pound increase ot 
pressure. 

For Parsons and Westinghouse-Parsons turbines the following corrections were used: 

Superheat (300-1000 kw.) 10 per cent.; (1200-7500 kw.) 8 per cent, per 100 degrees F. 

Vacuum (300-1000 kw.) 4 per cent.; (1200-7500 kw.) 3 per cent, per inch. 

Pressure .1 per cent, per pound, 
and the following for Curtis, Rateau, and Zoelly turbines: 

Superheat, 8 per cent, per 100 degrees F. 

Vacuum (26-28 ins.) 7 per cent.; (28-29.5 ins.) 8 per cent, per inch. 

Pressure .1 per cent, per pound. 

t Referred to 30 inches barometer. 



3 8o 



THE STEAM TURBINE 



and give a steam consumption of 13.19 pounds per kilowatt- 
hour (about 8.74 pounds per i.h.p.-hour) with steam supplied 
to the engine initially dry saturated (no superheat), 194 pounds 
per square inch absolute pressure and exhausting from the 
turbine at 28.8 inches vacuum, referred to 30 inches barometer. 

STEAM CONSUMPTION OF RECIPROCATING ENGINES SHOWING EXCEPTION- 
ALLY HIGH ECONOMY. 





T3 
C 

I 
I 


to 

I 

Q 

i 

1 

CO 


•si 
J* 

a 
i 

> 


w 
J5 
< 

03 

1 

i 

ca 

<u 

CO 


r.p.m. 


Steam per 
Hour. 


Equiv. 
Steam 
Consump- 
tion per 
b.h.p. at 
Deg. 
superheat, 

28 Ins. 
Vac. and 
165 Lbs. 
Abs. Pres- 
sure. 




Engine. 


i.h.p. 


b.h.p. 


References. 


Rockwood-Wheelock . . 

Mcintosh & Seymour. 

Leavitt Pumping En- 
gine. 

Rice & Sargent 
(Phila.). 

Westinghouse (verti- 
cal). 


595 
1076 
576 

420 

5400 

3600 
2000 

7500 

2500 
282 

142 



20 


297 

O 

308 
92 



223 
141 

256 


25.4 
27.1 

27-3 
25.8 

273 

27.6 
25-5 

251 

28.1 



27.2 


174 
138 
191 

157 

200 

146.5 
171 

190 

203 
156 

196 


76.4 
99-6 
51.6 

102 

76 

86 
100 

80 

85 
208 

195 


1300 
12.76 
11.20 

956 

11.93 

11 78 
11.05 

11.96 

8.96 
i5-24t 

9-65 


14.61 
14.19 
12.59 

10.75 

13 12 
12.95 


14.62 
1389 
1303 

13.39 

1376 
1358 


F. W. Dean, Trans. 

A.S.M.E., 1895. 
F. W. Dean, Trans. 

A.S.M.E., 1898. 
E. F. Miller, Tech- 
nology Quarterly, 

Vol. IX. 
D. S. Jacobus, 

Trans. A.S.M.E., 

1904. 
Eng. Record, May 

28, 1904. 






(Boston). 








York). 
Moabit (Berlin) 




H45t 




gine) . 


I0.82J 




Power, 191 2. 
Power, 1912. 


engine) . 







Effect of Superheat, Vacuum, and Admission Pressure on the 
Economy of Reciprocating Engines. According to Professor 
Schroeter* the steam consumption of reciprocating steam 
engines is reduced about 6 per cent, for 50 degrees F. and about 
9 per cent, for 100 degrees F. of superheat. Parsons f has 
shown that in a triple-expansion engine the steam consumption 
can be reduced only .4 per cent, per inch with an increase of 

* Storm Bull., Journal of Western Society of Engineers, December, 1903. 
f Proc. Inst, of Naval Architects, April, 1908; Mechanical Engineer, May 1, 1908, 
and Die Turbine, July, 1905. 

% Probably World's records for steam engines. 



STEAM TURBINE ECONOMICS 381 

vacuum between the limits of 25 and 28 inches, and at a still 
higher vacuum there is practically no gain at all. Increased 
initial steam pressure reduces the steam consumption of recip- 
rocating engines .1 to .2 per cent, per pound per square inch. 

A comparison of the two tables shows that in large capacities 
steam turbines will give, for the same standard conditions, 
better economy than reciprocating engines.* It is shown that 
for sizes from 3000 to 9000 kilowatts the steam consumption of 
turbines is about 12 pounds per brake horsepower-hour at the 
assumed standard conditions of o degrees superheat, 28 inches 
vacuum, and 165 pounds per square inch absolute pressure; and 
that, operating at the same conditions, the steam consumption 
of the best designs of reciprocating engines is about 13 pounds. 

Economy of Small Reciprocating Engines and Turbines. 
Nearly all small high-speed reciprocating engines rapidly deteri- 
orate in economy, primarily because the valve leakage becomes 
excessive. Although an engine of this kind will meet the guar- 
antees of steam consumption in a shop test, it has been shown 
that very soon they require a much larger amount of steam.f 
Tests of seven high-speed engines of various types rated at 100 
to 200 horsepower conducted by Dean and Wood in 1907 show 
that the steam consumption of such engines after a comparatively 
short duration of service was found to vary from 49.4 to 60.5 
pounds per kilowatt-hour at full load. These rates are very high 
when compared with the economy of small De Laval and Curtis 
turbines as given in Figs. 89 and 128. Parsons stated in 1904 
that the full-load steam- consumption of turbine-generators of his 
design under the conditions of 100 degrees F. superheat, 27 inches 
vacuum, and 155 pounds per square inch absolute steam pres- 
sure was approximately 25 pounds per kilowatt-hour for one of 
100 kilowatts capacity, while that of the 200 and 500 kilowatt sizes 

* It must not, however, be overlooked that these standard conditions were 
selected in the first place for comparing the economy of steam turbines. It happens 
that the vacuum is taken a little higher than is usual in the operation of reciprocat- 
ing engines. 

t "Economy Tests of High Speed Engines," by F. W. Dean and A. C. Wood, 
Proc. American Soc. Mech. Engineers, June, 1908. 



382 THE STEAM TURBINE 

was respectively 22 and 20 pounds. He stated that the equiva- 
lent results with dry saturated steam and 28 inches vacuum 
would be about ten per cent, larger.* 

It has been shown by repeated tests that the steam consump- 
tion of these turbines is not materially increased when operated 
continuously for long periods. Weithammer f states that he 
made tests of a De Laval turbine-generator when new and after 
five years of service, and calculated the deterioration in economy 
to be not more than two per cent. ; and this lower efficiency was 
probably largely due to wear of the reduction gears. It would 
appear that the deterioration of Curtis turbines should be even 
less because of less erosion from steam at very high velocities 
and the absence of the reduction gears. It is stated that there 
are cases where De Laval blades have been so much worn as to 
require replacing in a year. J Such an experience is, however, 
unusual. 

POWER PLANT ECONOMICS. 

The following table prepared by Mr. H. G. Stott of New York 
is interesting in many of its items. Actual data were used to 
determine the values under the heads of "Maintenance" and 
"Operation." The first column is for a plant with compound 
condensing reciprocating engines operating without superheat, 
and in all cases the values have been suitably corrected to make 
the other columns directly comparable with the first. 

Mr. Stott advocates the use of an exhaust steam turbine to be 
operated by the exhaust steam from reciprocating engines. By 
increasing the pressure of the steam supplied a moderate amount 
as well as superheating it the output of a power plant of the type 
represented by the first column in the table can be doubled at a 
comparatively small cost for turbines and boilers. 

* Trans. Inst, of Electrical Engineers, May, 1904. 

f Die Dampfturbinen, page 104. 

t Lea and Meden, Transactions American Soc. Mechanical Engineers, Vol. 25. 



STEAM TURBINE ECONOMICS 



383 



DISTRIBUTION OF MAINTENANCE AND OPERATION. 
(Charges per Kilowatt- Hour.) 



Maintenance. 

1. Engine room, mechanical. . . . 

2. Boiler room or producer room 

3. Coal and ash handling appa- 

ratus 

4. Electrical apparatus 

Operation. 

5. Coal and ash handling labor. . 

6. Removal of ashes 

7. Dock rental 

8. Boiler room, labor 

9. Boiler room, oil, waste, etc. . . 

10. Coal 

11. Water 

12. Engine room, "mechanical" 

labor 

13. Lubrication 

14. Waste, etc 

15. " Electric " labor 

Relative cost of maintenance and 

operation 

Relative investment in per cent. . . . 







Recipro- 




Gas 


Recipro- 
cating 
Engines. 


Steam 

Turbines. 


Engines 

and 

Steam 

Turbines . 


Gas 

Engine 
Plant. 


Engines 

and 

Steam 

Turbines. 


2-57 


0.51 


i-54 


2-57 


i-54 


4.61 


4-3° 


3-25 


I- 15 


i-95 


0.58 


O.S4 


0.44 


0. 29 


0.29 


1. 12 


I. 12 


1. 12 


I. 12 


1. 12 


2.26 


2. II 


1.74 


1 -13 


i-i3 


1.06 


O.94 


0.80 


0.53 


o.53 


0.74 


O.74 


o.74 


0.74 


0.74 


7-i5 


6.68 


5-46 


1.79 


3-°3 


0.17 


0. 17 


0.17 


0. 17 


0.17 


61.30 


57-3o 


46.87 


26.31 


25-77 


7.14 


7.10 


5-46 


3-57 


2.14 


6.71 


i-35 


4-03 


6.71 


4-03 


1.77 


o-35 


1. 01 


1.77 


1.06 


0.30 


0.30 


0.30 


0.30 


0.30 


2.52 


2.52 


2.52 


2.52 


2.52 


100.00 


86.03 


75-72 


50.67 


46.32 


100.00 


82.50 


77.00 


100.00 


91.20 



That the steam turbine plant has an inherent economy of 20 
per cent, better than the best type of reciprocating engine installa- 
tion is shown by a comparison of the first and second columns. 

Prices of Steam Turbines. Fig. 193 shows by means of curves 
the price per kilowatt of the normal full load rating of turbine- 
generators operating condensing. The prices given are the 
averages of those given by a number of manufacturers at a time 
when the cost of foundry pig iron was about $20 per ton. It is 
estimated that the values given by the curves will be changed 
roughly about 2 per cent, for a variation of $1 in the price of 
foundry pig iron. 

Unless some such standard of values is given such results can 
be of little value a very short time after the curves are prepared. 



3^4 



THE STEAM TURBINE 



Non-condensing turbines cost about .5 per cent, less than con- 
densing machines. Prices of 2 5 -cycle and 60-cycle generators 
are usually about the same. Prices do not include charges for 
freight and erection, which in the eastern and middle western 
states are about $1 to $1.50 per kilowatt. 



Pi 30 

I 

20 



























\ 


























\ 


























\ 




























^ 


*--. 




90 t 






























ivw 


h (iJ 


•c.c 


ten. 


































































«^20o" 


mo 


Km 


• (A 


C.G 


en. 



























































50 100 150 200 250 300 

1,000 2,000 3,000 4,000 5,000 6,000 
Rated Full.Load Kw. 



Fig. 193. Curves of the Approximate Price of Steam Turbine-Generators per 
Kilowatt of the Rated Full Load Output of the Generator. 



Mr. W. C. Gottshall, who has very carefully investigated power 
plant economics, has collected the data on the following page, 
published in 1903, regarding the probable maximum and mini- 
mum costs per rated kilowatt installed of a power plant equip- 
ment of about 10,000 kilowatts capacity.* 

High-grade power stations of from 5000 to 10,000 kilowatts 
capacity with thoroughly modern equipments cost usually from 
$100 to $125 per kilowatt. In a few very large stations with 
high-grade equipment the cost has been about $60 per kilowatt 
installed; but for stations under 10,000 kilowatts' capacity the 
cost is rarely below $90 per kilowatt. 

A building of modern factory type of construction (one story 
— steel and glass) costs about $1 per square foot of floor space. 

* Gottshall, Street Railway Economics. 



STEAM TURBINE ECONOMICS 



38S 



COST OF A STEAM POWER HOUSE AND EQUIPMENT. 



Boilers and settings 

Stokers 

Economizers 

Coal conveyors and bunkers 

Ash conveyor 

Piping and covering 

Feed- water heater 

Feed pumps , . 

Engines or turbines*: 

Generators 

Condensers including pumpst 

Switchboard 

Power-house cables and conduits 

Incidentals (as concrete floor and traveling crane) . . 

Foundations for machinery}: 

Buildings 

Chimneys and flues 

Total cost including 10 per cent, for engineer- 
ing supervision and contingencies (nearly) . . 



Costs per Rated Kilowatt 
Installed. 



Maximum. 


Minimum. 


$17.00 


Sq.oo 


5.00 


2.50 


4-5° 


2.50 


6.00 


2.00 


1.50 


I .OO 


12.00 


4.00 


2.00 


1 .00 


1. 00 


1 .00 


32.00 


20.00 


21.00 


18.00 


10.00 


2.00 


4.00 


1.50 


6.00 


3.00 


3.00 


2.00 


3-5° 


•5° 


15.00 


8.00 


2.00 


1 .00 



$87.00 



* This item is about right for reciprocating engines and turbines in 1903, when these data were pub- 
lished. Fig. 206 shows, however, that the minimum cost for this item is about the same as the present 
cost of first-class turbines and generators. 

t The cost of condensers is not included in Gottshall's data. Prices given here are those given 
by J. R. Bibbins {Report American St. Ry. Assn., 1004) for a plant operating at 26 inches vacuum. 
He estimates that the cost of a plant for 28 inches vacuum is 60 per cent, greater. Bibbins' values 
may be tabulated as follows: 

Cost of Condensing Plant per Rated Kilowatt. 



Barometric condenser 

Surface condenser (including centrifugal lift 
pump, air cooler, single-cylinder dry vacuum 
pump, and centrifugal circulating pu*np;.. . . 

Surface condenser (including wet vacuum pump 
and centrifugal circulating pump) 

Ejector condenser 



Inches Vacuum. 



$6 to $7.50 



$7.50 to $10 

S7 .50 to Sio 
$2 to S2 .50 



$9.50 to Si 2 



$12 to Si 6 



S12 to S16 

$3 to S4 



$12 to $15 



Si 5 to $20 

Si 5 to $20 
S4 to S5 



X Engine and generator foundations cost from Si .00 to S3 .50 per kilowatt capacity. Foundations 
for turbine-generators cost as a rule about one -fourth as much, usually 30 to 40 cents per kilowatt 
capacity on fairly good sub-soii. 



386 THE STEAM TURBINE 

Itemized costs of the 8500-kilowatt power plant of the Fort 
Wayne and Wabash Valley Railway Company at Fort Wayne, 
Ind., are given by Bibbins as follows: 

Substation apparatus and buildings are, of course, not 
included. Drawings and a photograph of this station are shown 
in Figs. 199 and 200. 

The double-deck arrangement and the installation of baro- 
metric condensers designed for a moderate vacuum make the 
first cost of this station very low. 

Dollars per 
Kilowatt. 

Building: Including general concrete and steel work, coal bun- 
ker, smoke flue, condenser pit, coal storage pit, etc IO -97 

Boiler plant: Including boilers, superheaters, stokers, piping, 

pumps, heaters, settings, breechings, and tank *3-9 2 

Generating plant: Including turbines, generators, exciters, cables, 

switchboards, transformers, and ventilating ducts 30 .55 

Condenser plant: Including condensers, pumps, piping, free 

exhausts, water tunnels, and intake screen 3.98 

Coal- handling plant: Including gauntree crane, crusher motors, 

and track o . 94 

Erection, superintendence, and engineering 5-94 



Total, excluding property and siding 66. 25 

The costs are based upon the following assumptions: 

Per Cent. 

(a) Bond interest and taxes 7 

(b) Sinking fund, equivalent to 6.43 per cent, depreciation 4.2 

(c) Total fixed charges on capital cost . . 1 1 • 2 

Depreciation determined by summing the depreciation on the 
several parts of the plant as follows: building, 3 per cent.; 
boiler plant and coal-handling apparatus, 10 per cent.; condens- 
ing plant, 6 per cent.; generating plant, 7.5 per cent.; general 
average, 6.43 per cent. 

In calculating the cost of an electric power plant it is necessary 
to consider the probable life of the plant, so as to make correct 
yearly reductions for depreciation. Often this is more or less 
of guess-work on the part of the constructor or owner, and for 



STEAM TURBINE ECONOMICS 



3S7 



this reason the following table from a recent issue of Zeitschrift 
des Vereines deutscher Ingenieure is interesting. The figures 
given are those used by two English public corporations, two 
English engineers, and a series of figures taken from German 
technical publications. Conditions in America are somewhat 
different from those of Europe and the depreciation is usually 
somewhat greater. 



ESTIMATED YEARS OF LIFE. 



Buildings 

Boilers 

Steam engines 

Steam turbines 

Gas engines 

Water turbines 

Dynamos 

Storage batteries 

Transformers 

Switchboards 

Electric cables (conductors) 

Electric meters 

Arc lights 



Authority. 



Local 
Gov't 
Board. 



L. Canby 
Council. 



Robert 
Hammond. 



J. F. C. 
Snell. 



3° 
15-25 



20 

5-7 

15 

15 

12-15 

5 



5° 
20 
20 



20 
20 
20 
20 
12-50 
10 



60 
20 
20-25 



15 

20 

3° 
10 
10 



60 
20 



10 

20 

20-25 

15-60 



German 
Publi- 
cations. 



66 

15 
20 
20 

17 

22 

20-22 

10 



15 



Comparisons have been made by L. G. French of the cost 
of two sizes of turbine-generators with corresponding recipro- 
cating engine costs. He states that the cost of a 750-kilowatt 
turbine-generator with a surface condenser (operating vacuum 
not given) and including foundations and installation charges 
was S3 7 per kilowatt. A similar reciprocating engine plant cost 
$40 per kilowatt. A 1 500-kilowatt turbine-generator with a similar 
condenser equipment cost $30.20 per kilowatt, including founda- 
tions and installation, while a reciprocating engine equipped 
similarly cost $32.40 per kilowatt. 

* Orrok of the N. Y. Edison Co. states that the company has had 40 small 
steam turbines of the impulse type in service for five years with practically no 
expense for repairs. 



388 THE STEAM TURBINE 

It is generally believed by engineers who have done recent 
work in the equipment of large new steam power stations that it 
is not very probable that large reciprocating engines will ever again 
be installed to develop power for electrical distribution. One 
reason is that turbine-driven alternators are particularly adapt- 
able for parallel operation. The low first cost and operating 
expenses of turbine-generator units as well as the saving in the 
cost of foundations and floor space are also very important 
considerations. A manufacturer of very large sizes of both 
steam turbines and reciprocating engines has stated that a large 
power station if equipped with reciprocating engines instead of 
steam turbines would cost at least from 35 to 60 per cent, more 
than a turbine station of the same capacity. 

The following interesting tests of the power required to operate 
the auxiliary machinery * needed for a Curtis turbine were 
reported by the Turbine Committee of the National Electric 
Light Association in 1905. The data apply to the auxiliaries of 



* " The quantity of circulating water required for high- vacuum condensing 
plants must be increased from the old standard of from 25 to 30 pounds to from 40 
to 60 pounds of water per pound of steam condensed for moderate temperatures, 
and from 60 to 100 pounds of water per pound of steam when used at the higher 
temperatures common to cooling tower practice. In cases of excessive head or 
quantity of circulating water the bulk of the power required by auxiliaries is due to 
the circulating pump. The range of power for this purpose varies so widely that 
the older method of assuming a given type of plant requiring 5, 10, or 15 per cent. 
of the total steam consumption to drive auxiliaries is entirely in error without an 
accompanying statement defining conditions under which circulating water is 
pumped. The power to drive the air pump is dependent somewhat upon the 
vacuum, but particularly upon the air leakage into the condensing system. It was 
for some time assumed that the work of the air pump corresponded to removing 
the air which entered the boiler in solution in feed-water. As a matter of fact, 
handling the air in solution is the smallest portion of work done by an air pump, the 
leakage through piping, pipe joints, pores of castings, stuffing-boxes, etc., imposing 
the greatest duty, the total quantity of air to be handled ranging from ten or fifteen 
to thirty or forty times the air dissolved in ordinary water. The actual power to 
drive the air pump should in good practice be less than .00018 indicated horsepower 
in the air pump cylinder per pound of exhaust steam per hour. As the amount of 
power necessary to drive the air pump is a comparatively small portion of the total 
power for auxiliaries a slight error in this quantity will not largely affect the final 
result. " — C. C. Moore, Journal of Electricity, Power, and Gas, March, 1905. 



STEAM TURBINE ECONOMICS 



389 



one of the 5000-kilowatt turbine-generators of the Boston Edison 
Company. 





Test 1. 


Test 2. 


Test 3. 


Kilowatts on turbine 

Vacuum 


27I3- 
28.4 

29-53 

13-9 
69. 1 

24-3 

6.4 
8.6 


3410. 
28.7 
29-95 

[orsepower Usee 

2 3-7 
69.1 

23.2 

5-8 
9.2 


4758. 
28.6 


Barometer 


29.96 

I. 

27.4 
69.1 
23-8 

5-6 
9.8 


Boiler feed pump 


Circulating pump 


Dry vacuum pump 


Step bearing pump 


Wet vacuum pump 




Totals 


122.3 
3-4 
8.4 


131- 
2.9 

7-4 


135-7 

2 . T 

5-7 


Per cent, power of auxiliaries to power 
of turbine 

Per cent, water used by auxiliaries to 
that used bv turbine 





PRACTICAL DESIGNING OF POWER STATIONS. 

Special Fields for Steam Turbines and for Reciprocating Engines. 

Some space will be given here to the economic considerations 
entering into the design of a modern power station intended for 
electric distribution of power. There can be no doubt as to the 
status of the steam turbine in comparison with reciprocating 
engines for the generation of electrical energy. Practically all 
the recently designed power stations for electric services are 
equipped with steam turbine-generators, and in some of the 
older stations built originally for engine-driven units it is not 
unusual to see turbines installed to increase the capacity.* 
There is a marked contrast between a power station equipped 
with reciprocating engines and one that is turbine-driven. The 
large and heavy frames, ponderous moving parts, and the 
large generators of reciprocating engine plants cannot be made 
to compete successfully with the smaller, more compact, and 
cheaper turbine units. But, on the other hand, reciprocating 

* W. C. L. Eglin, Report National Electric Light Association, 1906. 



39° 



THE STEAM TURBINE 



engines similar to the Corliss type have a field which for a num- 
ber of years probably, the steam turbine cannot enter success- 
fully. For irregular loads, suddenly applied, like those of rolling 




Fig. 195. Comparative Floor Space Required for Curtis Turbines. 

mills and mine hoists the reciprocating engine has advantages 
over the steam turbine, except perhaps when the turbine is used 
in connection with an electric drive. Because reciprocating 



STEAM TURBINE ECONOMICS 391 

water pumps and air compressors are more efficient, at least up 
to the present time, than centrifugal pumps and compressors, 
reciprocating engines are invariably installed in waterworks and 
compressing plants. In cotton and woolen mills and shops, 
where the power is transmitted by belts, shafting, and ropes 
instead of by electrical methods, the reciprocating engine because 
of its slower speed is generally preferable. 

Stated in a few words, the steam turbine is unrivaled by steam 
reciprocating engines for driving apparatus which can be operated 
efficiently at a high speed so that a direct-connected unit can be 
made. 

Parallel operation of alternators is greatly facilitated when they 
are driven by turbines rather than by reciprocating engines. There 
are always difficulties when reciprocating motion is to be con- 
verted into synchronous motion. Besides the advantages of a 
uniform turning moment which makes possible such close speed 
regulation that it is possible to operate railway, power, and 
lighting circuits from one turbine, because of its high speed it, 
produces a more powerful regulating force without the use of a 
fly-wheel than that of any engine-driven units of the same capacity. 
Where a steam turbine is installed in a plant with piston engines 
or water-wheels its inertia or fly-wheel effect has a steadying 
effect on the whole system. As an example of this inertia effect 
it is stated * that a 3500-kilowatt Curtis turbine and generator 
has at the rated speed (750 r.p.m.) a " storage energy " of 
30,600,000 foot-pounds which is sufficient to enable the machine 
to carry at any load an additional load equal to the full rating 
for about .75 of a second with a drop in speed of only 3 
per cent, and without additional steam. This machine could 
carry a momentary increase of load of half the rating for 1.5 
seconds. 

Floor Space for Power Plants. " Compactness " expresses *well 
the primary requisite for the economical design of modern power 
stations. The small space occupied by a Curtis steam turbine 
compared with that required for a reciprocating (Corliss) engine 

* A. H. Kruesi, Proc. American Street and Inter urban Railway Association, 1907. 



392 



THE STEAM TURBINE 



of the same capacity is well shown by Fig. 195. Floor space 
occupied by Westinghouse turbine-generators is given by the 
curves in Fig. 196, showing the number of square feet occupied 
per kilowatt or per- brake horsepower. Comparisons of the 
space required for power units are of little value, however, unless 
the space for the condensing apparatus and auxiliary machinery 
is also considered. It is probably fair to assume that for the 

































020 










Square Feet Floor Space per BMP. or KW 




^ .018 




\ 








1000-3000 KW Capacity 








\ 


V 






Westinghouse Double Flow Turbine 
5000-10000 KW. Capacity 












\ 


\ 




Over all Dimensions and Rated 






, O.IZ 




\ 










" 
















a 
^ 






\e 


\ 


\ 






















«5 








^ 
























1 


f 














<>v 








KW 








0.04 






















BMP 














0.02 




£ 


* as 




St 






^ 












^ 

* 











g 


1 i 






B.h 


p. 


1 


1 










1 










1 

1 


1 


<l 




1 














1 



5 











| 

5 


| 





Fig 196. Floor Space Required for Westinghouse Turbines. 



conditions where a reciprocating engine would be operated at 26 
inches vacuum condensers for a turbine plant would be designed 
for 28 inches vacuum. Now the volume of steam at 26 inches 
vacuum is very nearly half that at 28 inches. When surface 
condensers are used, therefore, the very great increase in the size 
of the condenser equipment for turbine plants is very obvious. 
For this reason there has been a tendency in recent years to 
install barometric or the open type of jet condenser for steam 
turbines. 

A very recent installation of Westinghouse-Parsons turbines, 
barometric condensers, and Stirling boilers is illustrated in 



STEAM TURBINE ECONOMICS 



393 




Fig. 197. "Double-Deck" Design of Power House Equipped with Horizontal 

Turbines. 



Fig. 197.* The important features of this design are the placing 
of the turbines above the boilers and condensers, the use of 

* J. R. Bibbins, Ttans. Am. Inst. Elect. Eng., 1908. 



394 THE STEAM TURBINE 

barometric condensers, and the low total cost of power house and 
equipment. It is probably one of the most compact arrange- 
ments possible in a steam turbine plant consistent with high 
economy in operation. 

In this "double-deck" design the horizontal turbine and 
barometric condensers are at their best advantage as regards 
compactness and efficiency.* The connections between the 
turbines and the condensers are short and direct, which obviates 
the losses occurring where there are bends in these connections, 
and the cost of large exhaust piping is saved. The atmospheric 
relief valve of the turbine is placed between the floor girders, so 
that it was possible to make the distance between the floor level 
and the condenser head only 2.5 feet. For stations not at tide- 
water, turbine plants are usually operated at a moderate vacuum 
of between 27 and 28 inches. Barometric condensers are now 
being made to maintain this vacuum without the use of auxiliary 
dry-air pumps. 

With surface condensers by far the most compact arrangement 
is obtained by installing Curtis vertical turbines with a condenser 
base. By this arrangement a very direct connection between 
the turbine and the condenser is secured; but in places where 
there is likely to be trouble with leaky tubes most engineers will 
prefer a condenser separate from the turbine. 

Drawings showing the cross-section and plan of a design for 
horizontal turbines, Babcock & Wilcox boilers, and barometric 
condensers are shown in Fig. 199. An exterior view of the 
same station showing the coal-handling equipment is illustrated 
by Fig. 200. 

Plan and elevation of a power station with the turbines and 
boilers on approximately the same floor level are represented by 
the drawings in Figs. 201 and 202. The first of these figures is 
particularly interesting because it shows very clearly the arrange- 

* A similar "double-deck" arrangement has been proposed for power plants 
operated by horizontal gas engines and producers. In such a design, where pro- 
ducers, scrubbers, and all auxiliaries are placed on the second floor, it has been 
shown that the ground-floor area was only 2.25 square feet per kilowatt. 



STEAM TURBINE ECONOMICS 



395 




TrPic«L FLOOR STRUCTURE FT. tVATNE STATION 



ftOOH.P. BATTCBY 
BO'LCff R001 fLOOtt LgVCt.. 



800 hf» BATTERY 



J lb LCVCL. OF FOUNDATION- jj fr MIC, PlCg<5 



lL_ik 



LONGITUDINAL FLOOR ELEVATION 




TRANSVERSE FLOOR FLEVATION 

Fig. 199. Longitudinal and Transverse Sections of Power Station, 



39 6 



THE STEAM TURBINE 



merit of the auxiliaries in stations equipped with surface con- 
densers. Fig. 203 is intended to show particularly the piping 
arrangements for a typical power station having the turbines 
and auxiliary equipment in a room adjoining the boiler room. 




Fig. 200. View of the Power House Shown in Fig. 199. 



OILING SYSTEMS FOR STEAM TURBINES. 

A perfect oiling system is obviously a necessity for any 
machinery operating at a high speed. The efficiency of turbines 
of the Parsons type depends largely on the smallness of the radial 
clearances between the rotor and the casing. Now if there is 
any displacement of the rotor with respect to the casing, caused, 
for example, by the melting of the white metal in one of the main 
bearings, the blading might be entirely torn or "stripped," and 
the turbine would probably be out of service for several weeks. 
In Curtis turbines with vertical shafts, on the other hand, very 
serious results might occur if the flow of oil to the step bearing 
should be interrupted. 



STEAM TURBINE ECONOMICS 



397 




Fig. 201. Plan and Elevation of Turbine and Condensing Plant. 



39» 



THE STEAM TURBINE 







STEAM TURBINE ECONOMICS 



399 



S^ss 




cu 



£ 
'3 
W 



E 

i 

< 

c 
o 



PE, 



I i 



400 THE STEAM TURBINE 

There are two usual methods of lubrication for steam turbines; 
(i) the central system and (2) the single unit system. 

In the case of the central system an oil tank is placed at a high 
point in the building and the oil flows through pipes by gravity 
to the bearings of the turbine. By means of a "parallel" system 
of piping any number of turbines can be supplied from one oil 
tank. The oil leaving the bearings flows into a suitable filtering 
apparatus provided with cooling coils from which it is pumped 
back to the main supply tank. The chief objection to this system 
is the danger of a total shut-down of the oiling system caused by 
a poor joint or a broken pipe between the supply tank and the 
turbine bearings. 

The alternate system, in which each turbine has its own oil 
supply and pump, has the advantage in that it assists in reducing 
the risk of a total shut-down of the plant to a minimum, and if 
the oil is spoiled in one turbine, due to being mixed with water 
or being overheated, the entire supply of the station is not 
ruined. 

Until recently nearly all manufacturers of Parsons turbines 
supplied their machines with plunger reciprocating oil pumps. 
In this respect an innovation has been introduced in Westing- 
house turbines by the use of a rotary oil pump shown in Fig. 204. 
In the drawings shown here there are two sectional views of the 
pump. A worm gear on the turbine shaft transmits power to 
the pump by means of the gear wheel 10. The direction of 
rotation of the shaft and of the flow of oil is shown by arrows in 
the sections. The pump cylinder and its rotor are not con- 
centric, and metal strips, backed by springs, are inserted into 
slots in the rotor. These strips are forced out by the springs to 
touch the inside of the pump cylinder in every position, so as to 
form pockets into which the oil enters on one side and is dis- 
charged from the other side. Similar rotary pumps are very 
generally used for all kinds of engineering services. 

A suitable oiling system for a Curtis turbine (including the 
step bearing) is well illustrated diagrammatically by Fig. 205. 
A large storage tank, shown at the right-hand side of the figure, 



STEAM TURBINE ECONOMICS 



401 



is fitted with suitable straining devices and a cooling coil. It is 
usually located low enough to receive oil by gravity from all 
parts requiring lubrication. Oil from this tank flows to a pump 
from which it is discharged at a pressure about 25 per cent, 
greater than that required to sustain the weight of the shaft and 
wheels on the step bearing. A baffler in the form of an adjustable 




L^JJ-10 




/Hi rfll I ffh tft* 

mm 



Tr 



i6 n~r 



Fig. 204. Westinghouse Oil Pump. 



spiral inserted in the pipe leading to the step bearing serves to 
regulate the oil supply. Another line of piping is provided for 
oiling the upper parts of the turbine.* This line of piping is 
provided with a reducing valve and an air chamber partly filled 
with compressed air to maintain a constant pressure necessary 
for the hydraulic motor operating the valve mechanism. Drain 

* Oil pressure on the upper bearings is about 60 pounds per square inch. 



402 



THE STEAM TURBINE 




^/i£L/£:f K»LV£ 



=, GAUGE GLASS 

fa/? CHARGING* 
\A/B ' CHAMBER 

THREE WAY COCM 



CH£CK K4LlS£~ 



TO SPRUNG EQUALIZER . 
OR ACQ U/WULATOR <=fl[ 



e(/MR 



^rOXAG-£ 7-AAfH 




Fig. 205. Oiling System for a Curtis Turbine. 



STEAM TURBINE ECONOMICS 403 

pipes from the upper bearings and from the hydraulic motor dis- 
charge into a common receiver in which the streams are visible, 
so that the oil distribution can be always observed. 

At some point in the high-pressure system adjacent to the pump 
a device is usually installed to equalize the discharge of oil from 
the pump. Ordinarily Curtis turbines are provided with a small 
spring accumulator for this purpose, except for cases where 
weighted storage accumulators are to be installed. A storage 
accumulator is usually recommended for large power stations. 
It can be arranged so that it will normally remain full, but will 
discharge if the pressure fails, and start automatically auxiliary 
pumps. 

Piping for Superheated Steam. Much of the trouble resulting 
from the use of superheated steam is due not so much to want of 
strength as to the want of elasticity in the parts affected. These 
troubles are due particularly to the unceasing variations in 
temperature resulting from fluctuating loads rather than from 
high temperatures. As it is possible for water to exist in the 
liquid state in superheated steam, the variations in temperature 
may produce a spraying of highly heated surfaces, which greatly 
increases these difficulties. Changes in the design of pipe 
fittings, valves, boilers, and superheaters should be made to allow 
for this abnormal condition. It is desirable to use annealed steel 
castings in place of cast-iron for fittings and valve casings, and 
the use of copper for internal parts of valves and gaskets should 
be avoided. Low velocities in steam piping, which have become 
customary on account of the pulsating flow of reciprocating 
steam engines, are not suitable for superheated steam. Since 
flexibility is so important a consideration in piping for super- 
heated steam, it is necessary to use comparatively small sizes of 
pipes and fittings. 

In Curtis turbines, Kruesi states, a velocity of at least 140 feet 
per second (about 8500 feet per minute) is desirable for dry sat- 
urated steam. Now if the steam is superheated 100 degrees F. 
the volume is increased 15 per cent., but "the velocity in the 
pipes will be substantially the same on account of the reduction 



404 THE STEAM TURBINE 

in the steam consumption of the turbine." Although this 
statement is not quite accurate because the steam consumption of 
Curtis turbines is usually reduced only 8 to 10 per cent, per ioo 
degrees F. superheat, it is an important observation that the size 
of piping should not be increased in proportion to the increase in 
volume of the steam due to superheating. 



CHAPTER XV. 
STRESSES IN RINGS, DRUMS, AND DISKS. 

Design of a Bucket Band or Ring. A ring or band is one of the 
simplest means of fastening together a number of separate pieces 
attached like the blades of a turbine wheel to the circumference 
of a cylindrical surface. Such bands are always made a little 
wider than the blades, especially at the side where the steam 
enters, so that the edges of the blades may not be easily damaged 
in transportation and from insufficient axial clearances when the 
turbine is operated. 

These bands are very serviceable in taking care of loose buckets 
which otherwise would be troublesome. The band serves to 
bind the blades together as a whole, making the blades with weak 
attachments to the wheel as good as the strongest. The band 
assists in making a row of blades of uniform strength.* 

The design of such a ring revolving at high speeds should be 
determined by careful calculations; but the theory underlying 
the design of such a ring serves also for the design of turbine 
drums and disks. 

Centrifugal forces more than any other considerations deter- 
mine the design of a blade ring or band for strength. These 
forces produce, of course, tension and a resulting expansion of the 
ring — both of significant importance. 

The centrifugal force (CF) in any sector (W pounds) of a 

* In a Parsons type it cannot be assumed, however, that because the blades can 
be made stiffer by the use of a band or shroud ring it is possible to reduce radial 
clearances below the normal amount and at the same time reduce leakage around 
the blades. There is reason for believing that radial clearances should be increased 
for satisfactory operation when the "band " construction is used unless the relative 
expansion of the metals in the ring, blades, drum, and casing is very carefully 
adjusted. 

405 



406 



THE STEAM TURBINE 



freely rotating ring of radius r inches, velocity V feet per second, 
with an angle 6 subtended by the sector, is 

WV 2 

T> (3i) 



CF = 



g 



12 



where g is the acceleration due to gravity. 

This centrifugal force tending to expand the ring by increasing 
its circumferential dimensions 
sets up stresses which, for 
the purposes of calculation, 
may be represented by tan- 
gential forces at the ends of 
the sector. These forces are 
necessarily equal for equilib- 
rium and are shown as T and 
T in Fig. 208. If the breadth 
of the sector is represented 
by m inches and the radial 
thickness by n inches, then 
the area of the section over 
which this stress is distributed 
is mn square inches; and if 
S is the unit tensile stress in pounds per square inch, each tan- 
gential force is expressed by 

T = mnS. 

This force T on the section is tangential, and since the radial 
centrifugal force (CF) must be equilibrated by an equivalent 
radial force Td * or for equilibrium 

CF =T0f, 
WV 2 




Fig. 208. Forces in the Blade Band 
or Shroud Ring. 



= mnS0 



g 



12 



* This relation is obvious from the geometry of the figure. It is, of course, 
not quite accurate, but very nearly correct for small values of d, 

f It cannot be assumed that at the moment of rupture the stress will be dis- 
tributed between the two sections. The assumption made in the equations is, 
however, very much on the safe side. 



STRESSES IN RINGS, DRUMS, AND DISKS 407 

Now if z is the weight in pounds of a cubic inch of the material 
of the ring, the length of the sector (Fig. 208) is rd inches; then 

W = mnrflz, 

5™M! _ mnSg> 

r 

Q I2ZV * , s 

S- — • (3a) 

This equation shows that the unit stress in a blade ring or band 
depends only on the weight of the material and on the peripheral 
velocity. The last equation can also be expressed in another 
form, remembering that 

v = 3.i4i6dN ^ 
60 X 12 

where d is the diameter in inches to the central line of the ring 
and N is the number of revolutions per minute. Then if we 
make the approximation of i? = 10, we have 

s zdV 



g x 4320 



s= ^iL. (33) 

139,100 K66J 

Equations (32) and (33) are generally used for the design of 
shroudings and overhanging rims. When such rings are per- 
forated with small holes for the riveting of blades, bending and 
shear are produced. The stresses due to this bending and 
shear are, however, small and do not in practical cases often 
exceed 400 pounds per square inch. 

Sometimes rings called " segments" (Fig. 115) are put on the 
edge of wheel disks and the blades are attached to them. In a 
construction of this kind the ring must not only restrain the cen- 
trifugal force due to its own weight, but also part of that from 
the weight of the blades if they are not tightly fitted. 



408 THE STEAM TURBINE 

If the following symbols are assumed: 

r = radius to center line of blades, in inches, 
d = diameter to center line of blades, in inches, 
w = weight of blades in pounds per foot of length of the 
circumference measured to the center line of the 
blade ring, 
rd = length of a short segment of the blade ring to be calcu- 
lated, in inches, 
V = peripheral velocity of blades, in feet per second, 
W = weight of the blades, in pounds, of a segment r d inches 

i wr o# A 

long or — "- pounds, 

12 

then the centrifugal force at the blade ring due to the weight of 
the blades alone is 

c _ W V 2 _ wr ft flV 2 _wflV ^ 

S 12 

Then if T is the tangential force in the blade ring due to the 
weight of the blades, and S is the corresponding unit tensile 
stress in pounds per square inch, 

C = T d = mnS o and 

g 

s -*¥t, 
gmn 

or 

518,400 gmn 51,840 gmn 1,669,200 mn 

then the total stress S t due to the weight of the ring and of the 
blades is 

c zd 2 N 2 , wdX 

Jj^ = J — , 

139,100 1,669,200 mn 

St = T ^ ft , ftn f 12 zd + Z±J ' (35) 

1,009,200 \ mn / 



STRESSES IN RINGS, DRUMS, AND DISKS 409 

If a blade band or shroud is made in a solid ring and is shrunk 
on the outside of the blades, as is sometimes the case, then the 
elongation of the ring due to the centrifugal stresses must be 
allowed for. In other words, the ring must be made small 
enough so that there will be a tight fit at the highest speed that 
will ever be attained.* 

Design of Drums for the Rotors of Reaction Turbines. The 
blades of steam turbines are, as a rule, fastened to a cylindrical 
drum or to one or more disks. The drum construction is used 
where there is a large number of stages with a small drop of 
pressure between the successive stages and usually comparatively 
low peripheral speeds. Thus the rotor of a Parsons type is 
made up of a number of drums of different diameters, increasing 
in size toward the low-pressure end. The drum diameters are 
determined by the blade speed which is selected by the designer 
to give approximately the best efficiency for the velocity of the 
steam in the stages of each section of the rotor. 

Calculations to determine the thickness of a section of the drum 
are the same in principle as for a blade ring as explained in the 
preceding paragraphs. 

The thickness of the drum shell is most simply determined by 
making calculations in the following order: 

(1) Calculate the stress in the cylindrical shell of the drum 
due to its own weight by equation (33). This stress can be 
determined immediately because it is independent of all dimen- 
sions of the drum except the diameter of the shell at its center 
line. It is assumed, of course, that before the thickness of the 

* If 5 = elongation per inch of length, 

5 = the unit stress lbs. per sq. inch in the ring at the maximum speed 

attained, 
E = modulus of elasticity in lbs. per sq. inch, 

then s = —, and the total elongation of circumference is inches. 

This means then that the circumference of the ring must be made — - — inches 

smaller than if not subjected to centrifugal stress. A very common construction is, 
however, that of making the ring in segments of about 2 feet in length and riveting 
the blades to these segments. 



4 io THE STEAM TURBINE 

metal for the drum is to be determined, the blades have been 
designed so that their weight can be calculated. 

(2) Allowable unit tensile stress must be determined. In this 
connection the factors to be considered are the qualities of the 
material to be used (see pages 429 and 430) and the grade of work- 
manship that is available. In some shops in Germany where 
very expert workmen can be secured and the material is carefully 
selected and unusually good, a factor of safety as low as three is 
sometimes used. Manufacturers of De Laval turbine wheels 
make the limiting factor from four to five; but for average 
American practice a factor of safety of less than five should not 
be considered. If nickel steel is to be used of which the ultimate 
strength is say 120,000 pounds per square inch, with a factor of 
safety of five, the allowable total stress in the drum shell would be 
24,000 pounds per square inch. Now if the stress due to its 
own weight, of which the calculation has already been indicated, 
is still represented by the symbol S, and the total stress allowable 
by S t , then the permissible stress resulting from the weight of 
the blades S is 

S = St — S = 24,000 — s. 

(3 ) The thickness of the drum shell can now be calculated by 

equation (34). Since S is now determined and d , N, andw* are 
given by the dimensions required for the design of the blades, 
the thickness n can be easily calculated. 
Equation (34) can be written in the form 

wd 2 N 2 , <N 

mn = • (36) 

1,669,200 X S 

Since the weight of the blades has been calculated for only one 
row, the dimension m is the distance between the center lines of 
successive blade rows on the drum. 

* Blades made of bronze, zinc, copper, or similar alloys weigh about .30 pound 
per cubic inch, and steel weighs .28 pound per cubic inch. 



STRESSES IN RINGS, DRUMS, AND DISKS 411 

Example. The following data regarding the shell of a section 
of a turbine rotor are given by the drawings accompanying the 
blade design. 

Diameter at root of blades (approximately = d) 25 inches 

Diameter at center line of blades (d ) 30 inches 

Revolutions per minute 2000 

Weight of blades in one row, per foot (w) 5 pounds 

Weight of a cubic inch of material of shell 28 pound 

Distance between center lines of successive rows of blades of 

the drum 3 inches 

The stress in the shell (S) due to its own weight, by equation 
(30), is 

c .28(2 5 ) 2 (2000 ) 2 , . . 

S = ^^ — — = 5030 pounds per square inch. 

S = 24,000 — 5030 = 18,970 pounds per square inch. 

5(3o) 2 (2000 ) 2 . , • 

mn = — JVO ~ = .57 square inch. 

1,669,200 X 18,970 

But m = 3 inches; then 

n = -57 -*" 3 = .19 inch. 

The sections of the rotor are usually supported on disks attached 
to the shaft. In another paragraph relating to the design of 
disks, the strength of such forms will be discussed. It should be 
remembered that, compared with impulse turbines, the peripheral 
speed is always kept low.* Drums are almost always used for 
reaction turbines, and separate disks or wheels for impulse 
turbines. 

Fig. 209 is an exact copy of the shop drawing of the rotor of an 
Allis-Chalmers (Parsons type) turbine. It consists of a central 
cylinder upon which rings are fitted as shown. These rings are 
made of steel and are forged as a solid ring. The webs are 
formed by cutting away the superfluous material in the sides 
with a lathe. In this type of rotor the central cylinder must be 
made of sufficient strength to resist the usual torsional stresses 

* The peripheral velocity of drum types should not exceed 400 feet per second. 
Impulse wheels, however, are sometimes designed to operate at 1200 feet per 
second. 



412 



THE STEAM TURBINE 



in a " hollow " shaft. The construction of the drums of typical 
Westinghouse turbines is shown in Figs. 107, 109 and 184. 

In impulse turbines where all the expansion of the steam takes 
place in nozzles placed in diaphragms, or partitions between the 
stages, there is a large drop in pressure between any two stages, 
and therefore leakage of steam between the stages will be much 
greater than with the small pressure drop in the reaction type. 
The fewer number of stages in the impulse turbine necessarily 




gTTTT 



Fig. 209. Section of the Rotor of a Parsons Type of Reaction Turbine. 

increases the velocity of the steam passing through the blades 
and at the same time the most economical wheel speed. Within 
practical limits, wheel speed should always be increased with 
steam velocity in good designing. 

Stresses at Right Angles to Each Other. To determine the 
stress in flat disks a refinement in the calculations is sometimes 
necessary in order to obtain more accurate values than those 
secured in the preceding calculations for the stresses in rings and 
drums. If, for example, two forces R and T act at right angles 
to each other, theoretical conditions of elasticity show that the 
maximum stress or elongation is never quite equal to that due 
to either of the two forces if acting alone. In other words, an 
elongation in the direction of the line of action of the force R 
produces a contraction in the direction of the force T.* Thus 

* This phenomenon is easily observed in a piece of india-rubber. A force in 
one direction producing an elongation will produce also a contraction in the 
direction at right angles to the greatest elongation. 



STRESSES IN RINGS, DRUMS, AND DISKS 413 

if the elongation due to the force R is s r per unit of length we have 
the relation 

where S r is the stress in pounds per square inch and E is the 
modulus of elasticity of the material. The reduction (s nr ) of the 
dimension at right angles or normal to the direction of the force 
producing the elongation is proportional to the force itself and 
also, of course, to the stress. Then 

where k is a constant and has the value of .3 for metals of a 
homogeneous structure, such as are usually required for the 
manufacture of machines. 
The force T in the same way produces elongation 

& 

S < = E 

and a reduction at right angles (s nt ) (in direction opposite to the 
elongation due to Sr), 

Snt== Y' 

The net elongation in the direction of the force R is 

Sr - *nt = ~ ~ — = " (Sr ~ -3 St). 

Also the net elongation in the direction of the force T is 

St kS* 1 /c c , 

s t - s nr =- - y = e ( ' "" * 3 r) * 

When the two stresses at right angles are nearly equal, as in 
the case of the disk now under consideration, the elongation is, 
from the results above, only .7 of that resulting from either force 

* See Greene's Structural Mechanics, pages 7 and 184; Church's Mechanics of 
Engineering, p. 203. 



414 



THE STEAM TURBINE 



R+,dR 



acting alone. It follows also that when the stresses are nearly 
equal the stresses which are, of course, proportional to defor- 
mations are also only .7 of that calculated from only one of the 
forces. This effect of forces at right angles to each other will 
be applied in the discussion of the stresses in disks. 

Mathematical Treatment of Stresses in Disks. Fig. 210 shows 
a section of a turbine wheel cut out (1) by two radial planes 

making the angle with 
each other, and (2) by the 
cylindrical surfaces with 
radiuses of r and r + dr. 
The two other bounding 
surfaces are the sides of 
the disk. The thicknesses 
of the disk are t at the 
radius r, and t + dt at the 
radius r + dr. 

If this sector is rotated 
about the center it 
develops the centrifugal 
force (CF). Acting on the 
surfaces of the sector are 
also the forces R and R + dR in the radial direction and the forces 
T, T in tangential directions. The two tangential forces T, T 
form the angle 180 — d degrees with each other, and their result- 
ant is approximately TO when 6 is a small angle. We have, then, 

Forces acting outward = R + dR + CF. 
Forces acting inward = R + T#. 

If we call the unit stress in the radial direction S r and in tan- 
gential direction St, then at a section at radius r (if all the dimen- 
sions are in inches) the following relations result: 

R = r#tS r , 

R + dR = (r + dt) (t + dt) (S r + d$ r )d, 

T = tdrSt, 

Td =tdrdS t . 




Diagram of Disk Stresses. 






STRESSES IN RINGS, DRUMS, AND DISKS 415 

If V is the velocity in inches per second and w is the weight 
of a cubic inch (the specific weight), then the volume of the 
sector is very nearly tdrdr, and 

wV 2 
CF = Mr — • 

g 

For equilibrium, the sum of the forces acting outward equals the 
sum of those acting inward, or 

R + dR + CF = R + Td, or 

(r + dt) (t + dt) (S r + dS r ) + tddrwV2 = r0tS r + tdrdS t . 

Dividing through by 6 and neglecting infinitesimals of the second 
order, we have 

wtdrV 2 
r(tdS r + S r dt) + tdr (S r - &) + ™±1- = 0m ( 37) 

o 

This general equation is not suitable for calculations, but by 
assuming conditions of uniform strength or uniform thickness 
the form can be considerably simplified. 

Disk of Uniform Strength. If we assume, then, uniform strength 
in the disk, the stresses throughout are constant, and if S' is the 
stress at any point, then 

S' = S r = St = constant value 
and therefore 

dS' = o, and substituting these values in equation (37) 

nftS r +-tV*dr =0, 
g 

dt , wV 2 dr 

dt , w w ft/Wr , , . , 

\- — X = o, and by integrating, 

w oPr 2 
logt+g;X-j : + K*=o. 

* K is a constant of integration. 



416 



THE STEAM TURBINE 



Now when r = o, t = t , and K = — log t , so that 

w li?X* > 

log t + — X — + (- log t„) = o, 

log (Tj = -^ x — .-"=» 



2gS' 



WV2 



- = e 2 g s" 

wV 2 



t 



2gS' 



(38) 



in which t is the thickness of the required section at the radius r, 
t is the thickness at the center, and e is the base of Naperian or 
natural logarithms which is equal to 2.7183 and log 10 e = 0.43429. 
All the symbols in these equations (including 2g = 773 inches) 
are in inch units. 

If tj is the minimum thickness of the disk, then equation (38) 
can be written w( y l2 _ V2) 

t=V 2gS ' 1 (39) 

where Vj is the peripheral velocity at the radius corresponding 
to tj 'and V is the velocity corresponding to t as before. 

If the disk is not made of uniform strength throughout, then 
Vj is the velocity where the portion designed for uniform strength 
begins. 

Equations (38) and (39) are generally used by the designers 

of impulse turbines, and for 
the conditions of average prac- 
tice they are sufficiently accu- 
rate. 

Design of the Rim. An 
enlarged section or rim is usu- 
ally required at the circumfer- 
ence of a disk for the attachment of the blades. Stresses in this 
section require careful consideration. 

In Fig. 211, t x is the smallest thickness of the disk where it 
joins the rim (at the radius rj and t* is the thickness and b 2 the 

* The change from linear to angular velocity was made to make integration 
simpler. 




Fig. 211. Section of a Turbine Wheel. 



STRESSES IN RINGS, DRUMS, AND DISKS 417 

breadth of the rim of which the center of gravity is at the radius 
r 2 .* Blades attached to the rim produce by the centrifugal force 
due to their weight the stress S 2 in pounds per square inch. 
Besides this there is exerted on the section of the rim the stress 
due to the centrifugal force of its own weight and also the radial 
stress (S r ) in the disk exerted over the thickness t r The expan- 
sion due to these forces acting on the rim must, for equilibrium, 
be equal to the expansion of the section of the disk where it joins 
the rim. The sum of the radial forces F r acting on the rim per 
inch of length may be stated then as 

wV 2 bt t 
Fr = S 2 t 2 + ^Ll^J _„ Srti> (4o) 

in which w is the weight of a cubic inch of the material of the 
rim, V 2 is the velocity at the radius r 2 in inches per second. 

Radial expansion of the rim (^ 2 ) is expressed by the following 
form if a is the area of the rim section in square inches and E 
is the modulus of elasticity in pounds per square inch, J 

F?-r 2 

* Because the contraction of the cross-section due to stresses at right angles 
(page 412), has been neglected in the derivation, equation (38) should not be used 
for values of allowable unit stress less than 15,000 pounds per square inch, as it gives 
thicknesses at the center, for low stresses, which are sometimes considerably too 
large. Practical designers who are required to use unusually low stresses for disks 
will find a suitable discussion in Jude's The Theory of the Steam Turbine, pages 
188 to 204. wV 2 

f Centrifugal force due to a weight of a cubic inch at r 2 is — , which becomes 

wV*b,t n „ , . . £ r 2 

- — - — — when multiplied by the area of the rim section. 

t It is easily shown that the tensile stress in a thin cylinder is 

F r r 

S = — = sE, 
a 

where s is the elongation per unit of length and a is the area of the section. Then 
the total elongation of the circumference (^t) is 

. 2 xF r r 2 2 

h = 



Ea 

and the radial elongation (X) is j 77 2 

3 — -L — 2 

2tt Ea 



4i8 THE STEAM TURBINE 

Since the radial and tangential stresses in the disk have been 
made equal in the original assumptions, the unit elongations in 
every direction must be equal, so that the linear expansion in the 
length r is 

^=5-5-8^, (42) 

where k is the coefficient of the contraction of the cross-section 
for stresses at right angles (see page 413). 

For conditions of equilibrium obviously X 2 = X 1} and substi- 
tuting equation (40) in (41) and equating to (42) we have 

_LL (s ,,, + ^_s,,,) = fl^W 

Usually the percentage error from writing r 1 for r 2 is very 
small, so that we have in simpler form, 



lLUt 2 + ZX^L* _ Srtl ) = (1- k) S r r„ 



2 



from which either b 2 or t 2 can be solved. In most cases, how- 
ever, t 2 is determined by the blade dimensions, so that b 2 is 
expressed thus: 

i^-s, sJ»-s, 

\ =-— -* r, = ^-± r,. (43) 

^ Vi 2_ ( l_ k)Sr ^ Vl 2 -. 7 S r 

In this equation- the stresses are in pounds per square inch, 
V t and g are in inches per second, t lf t 2 , r 1? and b 2 are in inches. 
Minimum Thickness of the Disk. The thickness of large disks 
at the smallest section is not determined by the allowable stress 
but by the requirements for safe transportation and by the liability 
of thin disks to become distorted and unstable in balance. Disks 
about 5 feet in diameter should have a minimum thickness of 
from .4 to .6 inch, depending on the quality of the material and the 






STRESSES IN RINGS, DRUMS, AND DISKS 419 

speed for which they are to be used; and for disks 10 feet in diam- 
eter the minimum thickness should be from .7 to 1.25 inches.* 

The breadth of the rim (b 2 ) calculated by equation (43) is 
the maximum value allowable, but the breadth can be made, of 
course, less than that calculated. There will be a smaller radial 
force at the rim of the disk than is necessary to produce the uni- 
form radial stress S r , and the disk will not be one of uniform 
strength. The stress at the center will be reduced very much 
less than that at the smallest section. 

If now equation (43) is used to calculate the minimum thickness 
t 1} with an assumed value for b 2 suitable for the design, negative 
values may be obtained. In this case a smaller value of S r 
must be used in the calculation. Limits for S r can be easily 
determined by putting S 2 = o in (43); then 

wV 2 



g(l-k) 

which is the tangential stress in a freely rotating ring or is the 
usual "fly-wheel" formula when k= o. 

Practical Example. Design of the Rim of a Disk Wheel. 

A disk wheel 50 inches in diameter is to be designed for an im- 
pulse turbine to operate at 3000 r.p.m. The minimum thickness 
(t x ) is .4 inch, and nickel steel is to be used with an allowable 
stress of 28,000 pounds per square inch, which weighs .28 pound 
per cubic inch (w). Approximately the radius (r x ) at the inner 
edge of the rim is 25 inches, so that V is 7860 inches per second 
(about 450 miles per hour). The wheel is to carry two rows 
of blades, so that the thickness of the rim must be made about 
3.5 inches. The weight of these blades is equivalent to a solid 
ring of steel around the rim .3 inch thick, f The weight of the 

* Minimum thickness for a wheel 3 feet in diameter is about .25 inch. An 
approximate rule for the minimum thickness of disks is 

i m i n = .008 d to .01 d, 
where d is the diameter in inches. 

t Centrifugal force of the blades on a wheel is probably most simply determined 
by this method of calculating from a drawing showing the dimensions of the blades 
the thickness of a solid band or ring of the same weight. 



420 THE STEAM -TURBINE 

blades per square inch of the rim surface is .3 X .28 = .084 
pound, and the stress S 2 per square inch due to this weight is 
(take g = 386 inches per second) 

S 2 = '— 7 X -^ - = 538 pounds per square inch. 

3 86 2 5 

Substituting these values in (43), 

28,000 x — - 538 

b 2 = Q , Q , , 2 3 ' 5 X 25 = 1.5 inches. 

.28 X (7860) 2 

^ '- - .7 X 28,000 

386 

The thickness of the section at the center (t ) is calculated by 
(38), using the same allowable stress as before for S': 

-.28( 7 86o) 2 

t = t e 772 x 28,000^ 

t = te' 80 = 2 .2 3 t, 

t = .4 X 2.23 = .89 inch. 

The expansion of the radius due to the allowable stresses in the 
disk can be calculated by (42), taking E = 30,000,000 pounds 
per square inch and k = .3, 

7 x 28,000 x 25 = qi6 

30,000,0000 

and the expansion of the diameter is .032 inch. 

If reaction turbines are to be operated at higher peripheral 
speeds than 350 feet per second, the stresses due to the cen- 
trifugal forces are too large to use a free drum construction, so 
that the drum must be strengthened with spokes or flat disks. 
It is considered better practice, however, to divide a drum into 
short sections, and calculate each section by the method explained 
here for disk wheels by the use of equations (38) to (42). The 
Allis-Chalmers Company uses this method for the low-pressure 
stages of its latest designs as shown in Fig. 209, although the per- 
ipheral speed of this section of the drum is usually less than 250 
feet per second. 



STRESSES IN RINGS, DRUMS, AND DISKS 



421 



Practical Example. Design of a Wheel Disk without a Hole. 

Stresses in disks are difficult because the areas over which the 
forces are distributed are not readily determined; besides, the 
forces are not uniformly distributed over any one of the areas 
to be considered. The stresses in a disk are calculated usually 
by determining the force acting on the " boundary" areas of a 
circular sector imagined cut out of the disk. Such a sector is 
shown in Fig. 212. The radius is r inches, the elementary radial 
thickness is dr, t is the thickness of the sector (measured parallel 




Fig. 212. Forces in a Sector of a Wheel Disk. 



to the axis of the shaft), and is the angle subtended at the 
center by this sector. The centrifugal forces cause tangential 
and radial stresses. If we imagine the disk made up of a series 
of concentric rings, laid side by side and touching, the tangential 
forces tend to break the rings in the line of the tangent, and the 
purely radial forces, on the other hand, will tend, as it were, to 
break out pieces which would be carried away in a radial 
direction. In Fig. 212 the tangential and radial forces are 
shown more simply than in Fig. 210 in the directions to equili- 
brate the centrifugal force CF. In other words, the tangential 
and radial forces shown are those balancing the centrifugal 
forces. 



422 THE STEAM TURBINE 

An actual design of a 50-inch plain disk of forged steel without 
a hole at the center for a Riedler-Stumpf turbine is shown in 
Fig. 213, and the following paragraphs show how the calculations 
were made. Diameter of the disk (d t ) is 46 inches (measured 
inside the rim, which is 2 inches wide). Smallest section of the 
disk is taken as .5 inch (see page 418). Speed is 4000 revolutions 
per minute. The allowable unit stress is 20,000 pounds per square 
inch, and the disk is designed for uniform strength. Weight of the 
blades is .09 pound per inch of the circumference, producing a 
centrifugal force of 1840 pounds per square inch at the smallest 
section of the disk. Now it has been shown* that in a flat disk 
the stress at the edge due to an external centrifugal load (like 
blades and shrouds) is superposable by simple addition to the 
stresses (both radial and tangential) in the disk due to its own 
rotation. 

The rim was calculated as in the previous example, and the 
thickness (t 2 in Fig. 211) was determined by the width required 
for the blades. From the inner edge of the rim the disk was 
given a constant thickness of .5 inch till the tangential stress 
alone as calculated by equation (33) exceeded the allowable 
limit, t 

* Jude, The Theory of the Steam Turbine, page 198. 

f It will be observed that these approximations are very much on the safe side 
because of the effect of " forces at right angles " (see page 412). It is probable, 
however, that whatever the form of the disk (if not abnormally irregular), the 
stresses at the center are slightly higher than the peripheral stresses. In case of 
undue racing due to the failure of the governing apparatus or other cause, a disk 
designed for uniform strength will fly to pieces from the center. A De Laval 
wheel without the usual "safety groove" near the rim when tested to destruction 
broke up entirely and projected large pieces through a cast-steel casing two inches 
thick. When, however, the customary groove was cut just inside the rim, only 
pieces of the rim were broken off when an excessive speed was reached, and no 
external damage was done. 

It is stated by Jude that the metal left between the "safety grooves" of a De 
Laval wheel is "only sufficient to carry the traction load of the vanes." From 
this fact the minimum thickness of De Laval disks can be easily calculated, 
as it is generally stated that the factor of safety at the groove is 5, and the section 
before the groove is cut is two-fifths larger. Allowable unit stress is probably 
taken at about 30,000 pounds per square inch. 



STRESSES IN RINGS, DRUMS, AND DISKS 



423 



For some distance from the rim toward the center we have the 
case of a flat disk. Now in a disk of constant thickness without 
a hole at the center both the radial and tangential stresses increase 
from the rim toward the center. At the outer edge of such a disk 
the radial stress is only that due to the cen- ^ 
trifugal force of the blades and rim, while the 
tangential stress is of considerable magni- 
tude and is always greater than the radial 
stress, except at the center, where they are 
equal.* Because both radial and tangential 
stresses at every section are approximately 
increased 1840 pounds per square inch by the 
centrifugal force due to the blades, the net 
allowable stress is 18,160 pounds per square 
inch. The point where the "increasing" sec- 
tion begins is determined then by the following 
calculations — substituting in equation (33): 



di 



18,160 X 139,100 
.28 X (4000 ) 2 
d t = 23.8 inches, or r x ■■ 



565, 
1 1.9 inches. 



Beyond this point toward the center of the 
disk the section has been made of uniform 
strength as calculated by equation (38). The 
calculation of the thickness where the diameter 
is 10 inches (r = 5 inches) is given by equa- __ 
tion (39), 

t 



■28 X 21.2 (io) 8 

t x e 7 2 3 x 18,160 



Fto. 213. Design of a 
Wheel Disk without 
a Hole at the Center. 



in which t x = .5 inch and e = 2.7183, then, 

t = V 45 

t = 1.56 t t = .78 inch. 

In the same way the thickness can be calculated for enough points 
to determine the profile. The section shown in Fig. 213 is a 

* Jude, The Theory of the Steam Turbine, page 200. 



424 THE STEAM TURBINE 

typical "flat" disk. To facilitate the forging of such disks the 
profile is not made exactly as calculated but is gradually tapering 
from the smallest section to the center. The increase in the 
thickness from the rim to the center is very small compared with 
many designs, approximating a " concavo-convex " form (Fig. 
214). It is argued by some designers that the treatment of the 
" concavo-convex " forms is entirely wrong and that most prob- 
ably it is not possible for the stresses all along the central plane 
to be either equal to or less than those at the rim by merely satis- 
fying equation (38), and that the metal in the bulging part of such 
disks has little influence in modifying the stresses in the central 
plane. Some of the best authorities agree that it seems reason- 
able that whatever the form of the profile of a disk the " stresses 




1 

Fig. 214. Typical Solid Disk without a Hole at the Center. 

in and about the central plane do not differ greatly from those 
in a 1 flat disk running at the same speed."* 

A typical solid disk without a hole for bolts or for the passage of 
a shaft through it is shown in Fig. 214. It was designed for a 
very much higher speed than the one in Fig. 213, so that it has 
a bulging from near the center. This design shows an ingenious 
method for the attachment of the body of the disk to the shaft. 
It will be observed that the disk is made with a very small sec- 
tion near the rim ; so that the stress there far exceeds that any- 
where else. If the wheel breaks it will rupture first at this 
smallest section and the rim and blades will be torn off. When 
these parts are gone the centrifugal force will be so much reduced 
on the part of the wheel remaining that there can be still a very 

* Jude, The Theory of the Steam Turbine, page 204. 



STRESSES IN RINGS, DRUMS, AND DISKS 425 

great increase in speed without further damage. This disk is 
designed for a factor of safety of about five at the smallest section, 
and about seven at every other section. 

In designing a disk for high speeds, obviously a section that 
gives approximately uniform strength from the rim to the center 
is desirable. Experience in such calculations has shown that a 
disk of the shape shown in Fig. 211 fulfills approximately these 
conditions.* This disk was designed for a speed of 20,000 
revolutions per minute. There is a centrifugal force of about 
.2 pound per inch on the outside of the rim, due to the weight of 
the blades which are of the irregular shape shown in Fig. 64. 

Disks with Holes in the Center. Up to this point in the dis- 
cussion of stresses in disks only designs similar to Figs. 213 and 
214, without a hole, have been considered. When, however, a 
hole is made near the center of a disk the stresses are greatly 
increased. There are no very reliable methods for determining 
the stresses in disks of arbitrary shapes with central holes. 
According to De Laval, any methods for " taking into account 
the hub influences in the calculation are only rough approxi- 
mations" to the actual conditions. It can be shown theoretically 
that a mere pin-hole at the center of a disk makes the tan- 
gential stress S t at the hole twice that in a disk without a hole. 
Indeed a small flaw near the center of a disk may seriously 
affect the magnitude of the stresses. For this reason, steel ingots 
with any traces of "piping" must not be used for forged disks 
to be operated at high speeds. For thick disks of the typical 
De Laval shape when perforated, the exact solution is appar- 
ently indeterminate. Methods of calculation for such irregular 
sections have been proposed which depend on the determination 
of the mean stresses of the whole section. Results from such 
methods are, however, of no value at all, as it is known that the 
maximum stresses are often twice the calculated mean stress. 

* Besides blow-holes and piping in ingots for drop forgings, most makers put 
holes into the disks for the attachment of tools for removing the disks from the 
shafts, and for balancing weights. Very few disks are made that do not have some 
holes. 



426 



THE STEAM TURBINE 



The fact remains, however, that disks for turbines are very 
commonly made with holes in the center for the shaft, and other 
holes besides are often made for the attachment of tools for forcing 
the disk from the shaft when the wheel is to be removed. Stress 
distribution near a central hole of a nearly flat disk can be approx- 
imately calculated if a disk of comparatively smaller diameter is 
imagined cut from its center, and this small disk is then assumed 
to be of constant thickness and subjected to a radial stress at its 

rim equal to the uniform 
stress in the large disk if it 
had no hole. The stresses 
in this small disk with the 
hole can be calculated with 
some degree of accuracy 
from equation (37) by put- 
ting dt = o, since t has been 
assumed constant in this 
small disk.* Tangential 
stresses calculated in this 
way for a disk 10 inches in 
diameter with a hole 1 inch 
in diameter are shown in 
Fig. 215. Radial stress is, 
of course, zero at the center so that it is not important. The 
large disk (Fig. 213) was designed to make the combined unit 
stress in the section 20,000 pounds per square inch, and it is 
assumed, therefore, that the radial stress on the outside of the 

* Simplified formulas for a disk of constant thickness are given by 
Eyerman, Die Damp/turbine, pages 88-90 ; Stodola, Die Dampfturbinen, pages 
160-161. 

The algebraic work involved in obtaining equations suitable for calculations is 
laborious and complicated. Because these equations are not used directly for other 
calculations they are not given here. This chapter on stresses is not intended to be 
an exhaustive treatment, mathematically, and the practical designer wishing to use 
the minimum factors of safety should carefully study the graphical solutions given 
by Stodola; but he should remember that these methods referred to are only 
approximations and in a great measure are justified only because they have stood 
the test when applied in practice. 




Fig. 215. Variation of Stress in a Disk 
caused by a Hole at the Center. 



STRESSES IN RINGS, DRUMS, AND DISKS 427 

small disk has this value. The curve shows that the maximum 
stress at the hole is 40,000 pounds per square inch and that the 
stress is rapidly reduced as the distance from the edge of the hole 
increases till it reaches the constant value of 20,000 pounds per 
square inch, for which the wheel disk was designed. It should be 
observed, therefore, that the stress at the edge of a hole at the 
center of a disk is twice that at some distance away from the hole.* 
It should be carefully noted, however, that this discussion applies 
only to holes at the center of a disk. Holes near the rim such 
as are often made for balancing the disk or as a safety device 
so that the rim will break first in case of excessive speed, 
would be allowed for in practice merely by the reduction of the 
section. 

It is, however, a good practice to make the section at the hub 
of a disk with a hole at the center of sufficient size to withstand 
the greatest stress that may come to bear at the normal speed. 
Fig. 216 shows how the disk in Fig. 214 should be modified that 
it may be put on a 4- inch shaft. The thickness (z) of the hub 
will be determined in the usual way as discussed in books on 
machine design. Only its length (t ) concerns this discussion. 
Eyerman-j- and Stodola give elaborate graphic methods for this 
determination, but they will not be taken up here, as they are of 
no general interest. For most practical purposes it is satisfactory 
to make use of the results shown in Fig. 215 and make the length 
of the boss (t ) twice the thickness at the same section for a 
disk without a hole. Instead of reducing the section abruptly 
in proportion to the reduction in the stress the use of a fillet 
(see curve ab in Fig. 216) of very gentle curvature gives by far 
the best construction. { 

Because the distribution of stress is changed when a hole is 
made in the center of a flat disk, the section where the radial 

* This has been shown by a mathematical demonstration and the development 
of suitable formulas by Grubler, Zeit. Verein deutscher Ingenieure, 1897, page 
860 ; Kirsch, Zeit. Verein deutscher Ingenieure, 1897, page 798. 

f W. Eyerman, Die Damp/turbine, pages 86-98. 

J Stodola, Die Dampfturbinen, 3rd edition, page 164. 



428 



THE STEAM TURBINE 



stress equals the allowable limit must be calculated by a different 
method from that used for the disk without a hole at the center. 

To determine from the general theoretical equations for the 
stresses in disks the diameter where the radial stress in a flat 
disk with a hole at the center has a definite 
value is very laborious and almost impracti- r: " : 

cable; but the following approximate and more 
or less empirical formula for the radial stress 
in a disk of uniform thickness with a hole in 
the center can be used conveniently. It is 
practically the same as that given by Cree and 
Jude * except that it has been simplified by 
grouping constants and changing the units to 
correspond with those used in the other equa- 
tions in this chapter. 

If S r is the radial stress in the disk in pounds 
per square inch at any diameter d t inches, V is 
the velocity at the periphery of the disk in feet 
per second, D is the diameter of the disk in 
inches, and d is the diameter in inches of a hole 
at the center, then 



9oD 2 \ 



+ d 2 



dx 2 



D 2 d 2 \ , . 

-dTT (44) 



Now if this equation is to be solved to deter- 




mine d lf it can be written 



d i 4 + f^T Sr-D'-d'W-DM^O, (45) Fig. 2*6. Design 
V 4 V / of a Wheel Disk 

and putting B = (2°l£ S r - D 2 - d 2 ) and 
C = D 2 d 2 , then 



with a Hole at 
the Center. 



dl = v/-B ±v /B!_ c> 
2*4 

* Jude, The Theory of the Steam Turbine, page 204. 



(46) 



STRESSES IN RINGS, DRUMS, AND DISKS 429 

This last equation is easily solved after obtaining the values of 
B and C from the dimensions of the disk and the allowable unit 
stress. 

There are two values of d, because the radial stress increases 
to a maximum value and then decreases to zero at the edge of 
the hole. The larger value of d t is always taken to determine 
the design because between the two values of d t the radial stress 
has its maximum value. 

In the design for this example S r = 18,160 pounds per square 
inch, D = 46 inches, d = 4 inches, and V = 500 feet per second. 
The value of B is then 1358, C is 33,856, and d x is calculated to be 
36.5 inches. 

The section from the 36.5 inch diameter inward toward the 
center is made of uniform strength and is calculated by the use 
of equation (39) in the same way as in the preceding examples. 

Permissible Stresses and Suitable Materials. It is considered 
safe generally to use ordinary forged or rolled steel, for velocities 
not exceeding 600 feet per second; and for lower speeds than 
this limit wrought iron can even be used if it is of exceptionally 
good quality. For speeds from 600 to 1000 feet per second 
crucible cast steel can be used. 

Nickel steel is recommended for turbine disks by the Krupp 
Company of Essen, Germany. This nickel steel has an ulti- 
mate tensile strength of 125,000 pounds per square inch and 
12 per cent, elongation before rupture. The elastic limit is 
about 95,000 pounds per square inch. It is stated by the 
Krupp Company that they will produce a nickel steel of still 
higher tensile strength but only about 6 per cent, elongation. 
With some small forged pieces of this material an ultimate 
tensile strength of 285,000 pounds per square inch has been 
observed, with an elastic limit of nearly 225,000 pounds per 
square inch. All De Laval turbine wheels used in America are 
made in Sweden of forged nickel steel, which is rather high in 
carbon. 

Allowable working stresses must, of course, be left to the 
judgment of the designers. An engineer of the Krupp Com- 



430 THE STEAM TURBINE 

pany states that stresses in the same direction may be allowed 
in turbine disks as high as one-third of the elastic limit. 

Since the centrifugal force and therefore also the unit stress is 
proportional to the square of the velocity, if a factor of safety 
of 4 is allowed, the breaking speed of the wheel will be twice the 
normal speed, and the elastic limit of the material is only about 
1.5 times the normal speed. 

Excessive stresses at a hole are " dissipated" very materially if a 
dangerous stress is reached at the edge of the hole. Before rupture 
can occur there will be an excessive elongation of the material as 
soon as the elastic limit is reached at the highly stressed section. 

STRESSES IN REVOLVING DRUMS. 
As appears from an inspection of the various figures given in 
the text illustrating modern designs, a drum is generally made up 
of a number of sections. The end sections are generally integral 
with the shaft. (See Fig. H2g.) These end sections are gener- 
ally designed as discs of uniform stress. The intermediate blad- 
ing is generally carried on a ring bolted to these ends; similarly 
the impulse wheel does not differ very greatly from shapes gen- 
erally adopted for impulse turbines. 

CRITICAL SPEEDS OF LOADED SHAFTS. 
With the high speeds at which steam turbines are operated 
the centrifugal forces due to even a small eccentricity of the 
rotating masses produce vibrations, excessive stresses, and 
" springing " of shafts. As the result of the eccentric forces the 
shaft is bent farther out of line, so that the centrifugal forces 
and the amount of the eccentricity are increased until the stress 
set up in the shaft by the bending produces a force equal to the 
centrifugal force, and the center of gravity and " center of work " 
coincide. If W is the weight of the rotating mass in pounds, e is 
the " original " eccentricity of the shaft in inches, x is the eccen- 
tricity in inches at N revolutions per minute, P is the force 
applied to the shaft at the point of attachment of the disk which 
will bend the shaft 1 inch, within the elastic limit, C. F. is the 
centrifugal force of the rotating mass, and k is a constant, then 
n „ kWN 2 
gx 



STRESSES IN RINGS, DRUMS, AND DISKS 431 

The bending of the shaft at this speed is x — e, so that 

(x-e)P = — — (47) 

gx 

and e , N 

The increased eccentricity due to rotation is therefore pro- 
portional to the original eccentricity of the shaft and increases 

kWu 2 
with increasing values of u and hence also of N. When — — = 1 

Pg 

x becomes oo-, that is, the deflection becomes exceedingly large, 
unless prevented, and would break the shaft. 

It has been shown by Cree * that the critical speed N c (r.p.m.) 
of a shaft with some flexibility in the bearings carrying a con- 
centrated load of W pounds is 

»-^\/|' «»> 

where E is the modulus of elasticity in pounds per square inch, 
r is the radius of the shaft in inches, 1 is its length (between two 
bearings) in feet, and a and b are the distances from the load to 
the bearings, in feet. This formula f is to be used for only a 
single concentrated load like the single wheel of a De Laval 
turbine. When there are a number of wheels with possibly also 
a revolving field of a generator on the same shaft, the problem 
becomes very complicated if the loads are considered separately. 
Experience with such calculations has shown that for the cases 
occurring in practice % the critical speed can be determined by the 
following simple equation derived for the case of uniform loading: 

N c = 155,000 r 2 y^, (49) 

where W is the sum of the several loads on the shaft. 

* Proc. Physical Society (London), vol. XIX. 

f In this formula the weight of the shaft is not taken into account. The influ- 
ence of the weight of the shaft on the critical speed can be easily calculated, but in 
practical cases it may be neglected without appreciable error. 

X This applies particularly to the cases of Rateau, Parsons, and Curtis tur- 
bines and turbine-driven generators and pumps. 



CHAPTER XVI. 
GAS TURBINES. 

The development of the gas turbine, which should combine 
the high thermal efficiency of an internal combustion engine with 
the mechanical simplicity of the steam turbine, has occupied the 
attention of a number of able engineers from time to time but 
without unqualified success. Because of the severe conditions 
due to the very high temperatures of the gases after combustion, 
there are many difficulties in construction which in a large 
measure offset the otherwise simple mechanical construction. 

It may well be said that the designer of gas turbines is between 
"the two horns of a dilemma." If he tries to utilize the gases 
at the temperatures resulting from expansion in a single normal 
nozzle, the nozzles and blades will deteriorate very rapidly, and 
for the best efficiency the speed of rotation of the turbine must 
be made too high for utilization for general power purposes with- 
out the application of reducing gears; and, if on the other hand, 
he cools the gases by the injection of water or excess air into the 
combustion chamber to make the temperature of the gases suitable 
for the materials available for machine construction, the high 
thermal efficiency stated by the simplest laws of thermodynam- 
ics* is, of course, not attained. 

Since the gas turbine is certainly not yet out of the experi- 
mental stage,f although there are commercial applications, it is 
not out of place to give some space to its history. 

Probably the oldest form of gas turbine is the ancient propeller 

Ti — T 2 
* The thermodynamic efficiency of a heat engine is expressed by — = — , where 

T\ is the initial and T% is the final temperature of the cycle. By lowering the value 
of Ti, the efficiency is reduced in much greater proportion than the reduction in the 
temperature. 

t It is stated that quartz nozzles have been used successfully, but they are 
very expensive and fragile. 

432 



GAS TURBINES 



433 




Fig. 218. A Chimney Turnspit 
or "Smoke-Jack." 



mechanism, known as a "smoke-jack," which was used for 
operating the turnspit* of large open fireplaces. An illustra- 
tion of this " smoke-jack" is shown 
in Fig. 218, which is a copy of an 
old drawing published in Bishop 
Wilkin's Mathematical Magic in 1680. 
A similar apparatus is described by 
Cardan about 1550. This mechanism 
was placed in the chimney and was 
driven around by the ascending cur- 
rent of hot gases from the fire. Its 
motion was transmitted by gearing 
and belting to the spit on which the 
joint of meat was carried in front of 
the fire. The power of this "smoke- 
jack" can only be estimated by the 
work of the turnspit dog which it replaced. It must, therefore 
be rated at least one " dog-power. " 

The earliest attempt to construct a gas turbine on scientific 
principles was probably made by Stoltze of Charlottenburg, who 
received a patent for what he called a "hot-air" turbine in 1873. 
This apparatus consisted of two turbines on one shaft, one acting 
as an air compressor and the other as a power turbine. The 
function of one of these turbines was to draw in and compress 
the air to about 40 pounds per square inch absolute. Part of 
this compressed air was then passed through a combustion 
chamber or furnace, where it supplied the oxygen required for 
the combustion of the gas or oil fuel. Another part went 
through a heating chamber and was later mixed with the gases 
of combustion from the furnace. The mixture of gas and 
air was then expanded in the second turbine. The useful 
power developed by such a turbine is the difference between 
that developed by the gas turbine and that required to drive 

* Turnspit is the name usually applied to the dog which was used to turn, by 
means of a suitable mechanical contrivance, a spit or long iron bar, pointed at one 
end, used to hang up meat to be roasted. 



434 



THE STEAM TURBINE 



the turbine-compressor. A turbine designed to develop 260 
horsepower has been constructed on this plan, but it has not 
been commercially developed. It is very doubtful, if all other 
difficulties were overcome, whether this method of air injection 
could give nearly as good economy as water injection. (See 
page 436.) 

Some attention has been given to the development of the 
explosion gas turbine, of which a very simple form is shown in 
Fig. 219. It consists of a combustion chamber E, of which one 
end is closed by a large valve A opening inward, admitting air 
through the parts B, B and fuel through tubes F, F opening into 
the valve seat. The mixture of gas and air is ignited by electric 
sparks at I, and the products of combustion are discharged from 



rtW 




Fig. 219. A Simple Explosion Gas Turbine. 

the chamber through a small opening J leading into the nozzle N, 
where air, as shown by the arrows, is mixed with the gases to 
reduce their temperature before they reach the blades of the 
turbine wheel W opposite the nozzle. 

It is a well-established fact that when a mixture of gas and air 
is exploded there is first a sudden expansion and then, because 
of the combination of the hydrogen in the burned gases with 
the oxygen in the excess air to form water, a vacuum is produced. 
This phenomenon is applied in this apparatus to operate the 
valve A, which by the formation of a vacuum is drawn inward 
to admit another charge of gas and air. It is stated that in 
such a turbine the explosions will occur very rapidly — from 3500 
to 5000 per minute — so that there is a practically continuous 
discharge upon the wheel. The efficiency of an explosion motor 
of this kind is very low because of the lack of compression; but 



GAS TURBINES 



435 



its efficient development does not seem to be impossible. If in 
some way efficient combustion by explosion can be secured 
without compression, then a most economical power development 
could be attained with an explosion combustion chamber with 
the fuel and air valves operated automatically "by vacuum" and 
the injection of probably comparatively large quantities of 
water after combustion. Such an apparatus would be simple 




Oil Pump 

Fig. 220. Section of a Zoelly Explosion Gas Turbine. 

indeed compared, on the one hand, with the complicated com- 
bination of the steam boiler with external firing and the steam 
turbine, or, on the other hand, with the complex reciprocating 
gas engine. Fig. 220 illustrates a Zoelly explosion gas turbine. 
It consists essentially of an explosion chamber C, a turbine 
wheel W, water and oil pumps, and an air compressor. The 
pumps and compressor are of the reciprocating type and are 
driven by the main shaft by means of the worm gears A t and 
B ± . The valves regulating water, oil, and air admission and 
the ignition device ri, are operated by the gases, and steams are 



43 6 THE STEAM TURBINE 

expanded in the nozzle N and impinge upon a turbine wheel W 
of the De Laval type. Some of the heat remaining in the exhaust 
gases is absorbed by water coils R which serve to heat the injec- 
tion water. In the operation of this apparatus, air is admitted 
first into the explosion chamber and then the oil, as the air is 
supposed to act as a shield against back-firing. After the charge 
has been exploded and the maximum pressure has been reached, 
the cooling water is injected. 

The more successful gas turbines, however, are those operating 
by combustion at constant pressure. In this type the air and 
fuel (oil or gas) are delivered underpressure to a suitable com- 




Water- 

^Gas 

Fig. 221. Diagrammatic Illustration of the Combustion Chamber and Steam 
Coils of a Modern Gas Turbine. 

bustion chamber A in Fig. 221 which is maintained at a red 
heat, so that the combustion is continuous. The products of 
combustion are usually cooled by water which is injected into the 
nozzle as in the explosion type. The heat energy in the burned 
gases is converted into velocity in an expanding nozzle N and are 
discharged at a high velocity upon the blades R of the turbine 
wheel. Designers of this type of gas turbines have generally 
assumed that nozzles and wheels of the De Laval type are most 
suitable, and their energies are devoted at present to the pro- 
duction of a suitable combustion apparatus and a high efficiency 
rotary compressor. Fig. 222 shows a typical small gas turbine 
set up for a brake test. 



GAS TURBINES 



437 




Fig. 222. A Gas Turbine set up for a Brake Test. 



438 



THE STEAM TURBINE 



In practice the combustion chamber is lined with carbo- 
rundum, and to allow for expansion the carborundum is backed 
with sheets of asbestos to provide a soft and elastic packing. 
Exhaust gases are usually discharged over a coil boiler L, and 
the steam which is produced is also delivered upon the turbine 




Fig. 223. Arinengaud and Lemale's Gas Turbine. 



wheel by a separate nozzle M. When the turbine is in operation 
the lining becomes sufficiently hot to ignite the fuel as it is 
forced into the chamber. 

A gas turbine of this latter type designed by Arinengaud and 
Lemale of Paris is illustrated in Fig. 223. It is a machine 
developing 300 net horsepower at 4000 revolutions per minute. 



GAS TURBINES 



439 



A Rateau turbine-compressor shown direct connected to the gas 
turbine in Fig. 224 has been specially designed and built by 
Brown, Boveri & Co. of Baden, Switzerland, for use with this 
turbine. The compressor gives a mechanical efficiency as high 
as 65 to 70 per cent, and delivers 1 cubic foot of air per second 




Fig. 224. Arinengaud and Lemale's Gas Turbine Direct Connected to a 
Rateau Turbine-Compressor. 



at a pressure of from 6 to 7 atmospheres. Compressed air is 
used for starting, and a simple ignition device is used for firing 
the charge till the combustion chamber becomes sufficiently 
heated. M. Barbezat, who has now charge of the development 
of this turbine, states that the total efficiency is not as high as 
that of reciprocating gas engines; but no data are given. 



440 THE STEAM TURBINE 

Gas turbines have been applied practically for the propulsion 
of submarine torpedoes. Formerly some types of torpedoes 
received their motive power from a rotary motor like a turbine 
wheel, driven by compressed air. Recently gas turbines have 
been installed with an obvious gain in power and saving in weight. 
These gas turbines develop 120 horsepower at 1000 revolutions 
per minute. The expansion ratio of the nozzles is 8.4 and the 
weight per horsepower, without the compressor, is 1.3 pounds. 

It is obvious, then, that great progress has been made recently 
in the development of the gas turbine; and when the ratio of 
progress is compared with the time required to bring the recip- 
rocating gas engine to its present state of development, there is 
reason for hoping for greater accomplishments in the near future. 
The gas turbine question includes, however, a number of un- 
solved problems; but, on the other hand, the sources available 
for their solution are numerous. The development of these 
machines will permit the utilization for power of mixtures of air 
with coal gas, petroleum, or alcohol; and it will also make possible 
a combination of the explosion motor and the steam turbine for 
many purposes. 

The problem is laid plainly before the physicist, the engineer, 
and the machinist, and to bring about a satisfactory solution will 
doubtless require all their combined resources. 

Questions of Theory. The success of the steam turbine 
naturally directed the attention of engineers to the possibilities 
of the gas turbine with the expectation of combining the high 
thermal efficiency of the gas engine with the constructive advan- 
tages of the steam turbine. 

As explained in the preceding pages a gas turbine can be 
operated by either of two methods: 

(1 ) By combustion of the fuel in a chamber at constant pressure. 

(2) By an explosion method. 

Combustion at constant pressure seems to be the more practi- 
cable method and is the one generally adopted.* In the opera- 

* Theoretically the same efficiency should be secured with either of these two 
systems of combustion. Combustion at constant pressure is an adaptation of the 



GAS TURBINES 44* 

tion of this method gas and air are compressed in separate 
chambers or compressor tanks to a suitable pressure, usually 
about ioo pounds per square inch absolute. The gas and air 
are admitted through separate valves to the combustion chamber, 
where the gas is ignited and burned at constant pressure. Just 
as in a reciprocating gas engine, the air is provided to furnish the 
oxygen to support combustion. After combustion the burned 
gases escape through a suitable nozzle to impinge on the blades 
of the turbine wheel. On account of the extremely high temper- 
atures resulting from the combustion (about 2500 degrees F.), 
it is impracticable to design a gas turbine with more than one 
pressure stage and therefore only "nozzle types" can be used. 

Comparison of Losses in a Gas Turbine and in a Gas Engine. 
It is reasonable to assume that the radiation and cooling water 
losses will be about the same for a gas turbine as for a recipro- 
cating gas engine; and from a practical viewpoint the work 
required for the compression of gas and air is about the same 
for combustion at constant pressure as for explosion. After 
eliminating, therefore, the radiation, cooling water, and com- 
pression losses the same energy remains for utilization in each of 
these two prime movers. In gas engines from 20 to 25 per cent, 
of this energy is lost in the suction and exhaust resistances, engine 
friction, and the heat loss in the exhaust. Corresponding to these 
losses in the gas engine, there are in the gas turbine losses due 
to nozzle, blade, and disk friction, the heat in the exhaust, and 
bearing friction. The sum of these latter losses in a steam tur- 
bine would be about 40 per cent., and they will probably be not 
much different, in the total, in a gas turbine. It is argued in 
favor of the gas turbine that it is not impossible "to isolate the 

well-known Brayton cycle. It has been shown that exactly the same thermal 
efficiency can be secured by such combustion as in the ordinary explosion process 
if it is assumed that the specific heat of the gases is practically constant and that the 
final pressure after compression in the explosion motor is the same as the constant 
pressure of combustion in the Brayton cycle. It follows then that the ideal gas tur- 
bine will theoretically operate with the same fuel consumption per unit of power 
as the ideal four-cycle gas engine. (Cf. Lorenz, Zeit. Verein deutscher Ingenieure, 
1900, page 252.) 



442 THE STEAM TURBINE 

combustion chamber internally " so that no cooling water will 
be needed. 

The greatest practical difficulty in the way of the successful 
operation of gas turbines results from the high temperature of the 
gases at end of expansion. Nozzles can be cooled by water- 
jacketing, but the wheel blades are liable to rapid deterioration. 
The necessity for lowering the temperature of combustion is now 
generally recognized, and the hot water from the water jackets is 
sprayed into the compressed air supplied for combustion. By 
this means the temperature in the combustion chamber can be 
greatly reduced but at a considerable loss, however, in efficiency. 
It is making, in other words, the efficiency of the gas turbine 
approach the lower thermal efficiency of the steam turbine. 

It seems probable that the most promising field for the gas 
turbine will be found to be in the utilization of bituminous coals 
forming tar and asphaltum when used for making gas. Such 
coals cannot be used in the manufacture of gas for reciprocating 
gas engines, as the accumulation of tarry matter in the cylinders 
is particularly objectionable. In a gas turbine, however, the gas 
is burned under pressure, in an enclosed chamber where accumu- 
lations of foreign matter cause no serious difficulties. 

It cannot be expected that gas turbines can be commercially 
successful for general power purposes if a reciprocating compressor 
must be used in connection with them, because a gas turbine with 
a compressor of this type is quite as complicated as the recip- 
rocating gas engine. Compressors of the rotary type, on the other 
hand, have usually a very low efficiency; probably in most cases 
not more than 50 per cent. Decided progress is being made in 
the successful designing of compressors of the turbine type which 
will give from 60 to 70 per cent, efficiency. It is not difficult to 
understand how the net useful work of a gas turbine may be nil 
under conditions that are not unusually poor. For if the effi- 
ciency of the turbine is 60 per cent, and the theoretical work of 
compression is 40 per cent, of the output (which is not an absurd 
estimate), then with a compressor efficiency of only 40 per cent, 
the theoretical power absorbed by the compressor is 60 X .40, or 



GAS TURBINES 



443 



24 per cent, of the output, or the actual power delivered to the 
compressor is 24 -5- .40, or 60 per cent, of the output. And the 
compressor takes all the power the turbine can supply. It is 
obvious then that compressors with the usual low efficiencies of 
the rotary types are not worth considering. 

BRAYTON CYCLE CALCULATIONS FOR GAS TURBINES. 





2 K 
*l $ 

t/i 


H ° 


13 



1* . 


mpera- 
1 for a 
nly. 


Tature, 
for a 
nly. 


g i 


J c 


i! 

C y 




£ u ■ 

1) 
a tsc 


.0 g 

T3 t3 rt 


v ■ O 

H .3 


s. ° 

B.-B 3 


£ 
2 




•a fe 

l a 

JJ 



~£ 




aximum 
Pounds 
Inch Gau 


eat Adde 
per Poun 
ing Subst 


Dunds of 
Pound 0: 
Substanc( 


aximum 
ture, Deg 
Perfect G 


inal Te 
Fahrenhe 
Perfect G 




<4-l "^ 

fe 


elocity o 
Wheel, Fe 
ute. 




a 


X 


&. 


8 


fe 


rt 


> 


2 




Pi 


Q 


X 


k 


h 


r 


V 


e 


Case I: 


90 


1,000 





4,665 


2,435 


. 22 


115,3°° 


43 


Adiabatic com- 


90 


250 





i>5°5 


652 


•87 


71,680 


43 


pression. Per- 
fect machine 


19s 
195 


1,000 

250 






4,860 
1,710 


2,008 
546 


.27 
1.07 


131,200 
83,74° 


54 
54 


with no losses. 


495 


1,000 





5,*9° 


1,562 


•34 


147,600 


64 
64 




495 


250 





2,040 


434 


1.38 


98,300 


Case II: 


















Isothermal com- 


90 


1,000 





9, 2 39 


5,°39 


.08 


159,100 


93 


pression with 


90 


250 





1,965 


9i5 


•39 


79,57° 


72 


regenerator. 


195 


1,000 





7,376 


3^76 


. 11 


159,100 


9i 


Perfect ma- 


195 


250 





1,498 


448 


.61 


79,57° 


62 


chine with no 


495 


1,000 





6,087 


1,887 


•15 


159,100 


87 


losses. 


495 


250 





1,176 


126 


1.03 


79,57° 


49 


Case III: 


















Isothermal com- 


56 


333 





2,147 


1,200 


•74 


75.4°° 


27 


pression with 


82 


39 2 





2,398 


1,200 


.68 


84,800 


3° 


regenerator. 


61 


445 





2,867 


1,619 


•5o 


86,485 


3i 


Actual ma- 


98 


545 





3> 2 33 


1,619 


•47 


99,960 


35 


chine with as- 


33 


445 





3> J 79 


2,i39 


.40 


80,155 


28 


sumed losses. 


61 


569 





3,699 


2,i39 


.36 


96,775 


33 


Air excess. 


















Case IV: 


















Isothermal com- 


















pression with 


79 


625 


•37 


i,275 


600 


.67 


75,000 


13 


regenerator. 


47 


638 


.36 


1,680 


1,000 


•43 


75,000 


15 


Assumed losses . 


















Cooling water. 



















444 THE STEAM TURBINE 

THERMODYNAMIC THEORY OF THE GAS TURBINE. 

The elementary thermodynamics of the gas turbine involve 
apparently no new investigations. The problems are princi- 
pally mechanical and metallurgical. In the above table the 
efficiencies of various gas turbine cycles are given as calculated 
by Sanford A. Moss. In all cases it is assumed that the heat 
of combustion is developed at constant pressure and that the 
exhaust gases are discharged at constant pressure. Cases I and 
II refer to theoretically perfect engines, and cases III and IV to 
engines with probably normal losses. It is assumed for these 
latter cases that the turbine efficiency is 70 per cent, and of the 
compressor is 83 per cent.* The efficiency of the regenerator 
used for the cases of isothermal compression is taken to be 60 
per cent. These figures are certainly above the upper limits of 
possible results in practice. 

Thermodynamic efficiencies of gas turbines operating with 
combustion at constant pressure will now be discussed. The 
equations given are, in most cases, those relating to a perfect 
gas. 

The total heat H of a gas at constant pressure may be expressed 
by the following equation: 

H = c„T + RT = (c„ + R) T = c p T + constant, 

where T is the absolute temperature, c v and c v are respectively 
the mean specific heats of the gas at constant volume and con- 
stant pressure between zero temperature and T, and R is a con- 
stant varying for its value with the kind of gas. 

Fig. 225 represents the cycle of operations when a pound of a 
mixture of gas and air is compressed and later expanded in doing 
work. Adiabatic compression is assumed. One pound of the 
mixture is taken into the compressor cylinder at the temperature 
T and volume v and is compressed as represented by the adia- 
batic 3 to the temperature T 3 and volume v 3 . In the passage 
to the combustion chamber it will be assumed for simplicity in 

* Efficiencies of 60 per cent, for the turbine and not more than 70 per cent, for 
the compressor would probably be more reasonable. 



GAS TURBINES 



445 



the calculations that the temperature drops to the initial temper- 
ture T . If the total heat contents at the points o, 3, 3^ 4, and 
5 are represented by the corresponding symbols H , H 3 , H 3 ', H 4 , 




Fig. 225. Diagram of the Theoretical Action of a Gas Turbine and 
Air Compressor. 



and H 5 , the indicated work of compression is the area 0123, 
which will be represented in heat units by 



W, 



H 3 - H . 



It is assumed, however, that immediately after compression the 
temperature falls to T , so that the volume is reduced from v 3 to 
v 3 \ Now because the point 3' is on the isothermal 3', it is 
obvious that H = H 3 '. During combustion a quantity of heat 
Q x is added to the mixture, increasing the temperature to T 4 and 
the volume to v 4 ; or, in other words, 



Qi=H 4 



H. 



When the gaseous mixture is expanded the work W t is per- 
formed, which may be calculated after determining H 4 in the 
preceding equation; then 



446 THE STEAM TURBINE 

The heat lost by the exhaust gases in cooling from T 5 to T is 
H 5 — H ; and since this quantity is called Q 2 , we can write 

W, = (H 4 - H/) - (H, - H.) = Q t - Q 2 , 

and placing H for H 3 ' it is apparent that 

W, =H < -H 5 =Q 1 -Q 2 . 

The theoretical discharge velocity V in feet per second at 
the mouth of the expansion nozzle is calculated from the usual 
equation, 

V 2 



2g X 778 



= Q t -Q 2 =H 4 -H 5 =c p (T 4 -T 5 ). 



Using the same symbols as before for the indicated work of 
compression, the theoretical effective power of the turbine is 

W e = W t - W c = Q, - Q 2 - W c . 

If the mechanical efficiency of compression is x and the efficiency 
of the gas turbine, y, is determined in the same way as for a steam 
turbine, by constructing velocity triangles and calculating the 
nozzle, blade, and wheel friction losses for a single stage turbine, 
then the theoretical net power of the turbine is 

W 
W/ = (Q, - Q 2 ) y - -^- 

Since the heat consumed per pound of the mixture is Q lf the 
total efficiency z of the gas turbine apparatus is 



z = 



((Qi-ftfr-^+Qt. 



Efficiency of Gas Turbine with Water Injection. If m pounds 
of water are injected into the combustion chamber just before 
the expansion begins, an equal weight of steam is found, which 
it will be assumed is superheated to the temperature T/, which 
will now be also at the temperature of the mixture, lower, of 
course,, than T 4 . 



GAS TURBINES 447 

The temperature of the mixture of burned gases and steam 
T/ is calculated by solving the following equation: 

Cp (T, - T/) = m {q/ - q t + r/ + c/ (t/ - tJJ, 

where q/ is the heat of the liquid, r/ is the heat of vaporization, 
and t S4 is the temperature of saturated steam, — all at the 
corresponding pressure P/.* The other new symbols are q t -, 
which is the heat of the liquid at the injection temperature, and 
c/, which is the specific heat of superheated steam. In this 
equation t/ and t s4 are ordinary (not absolute) temperatures. 

The temperature T 5 ' is calculated for assumed adiabatic expan- 
sion by using the exponent k' calculated from the equation 
below : 

k / _ c Pl + mc p? 
c wi + mc w 

in which the subscript i refers to the specific heats of the mixture 
and the subscript 2 to specific heats of the steam. The temper- 
ature T/ is used to determine the value of Q 2 , which is the quan- 
tity of heat abstracted from the mixture to cool it from the 
condition at 5 to the condition at 0. It is calculated from the 
following equation: 

Q 2 = m ]c p ' (t/ -t S5 ) + r/ + q/ - q } + c P (t 5 ' -t ). 

In this equation t/ is the " ordinary" temperature corresponding 
to the absolute temperature T 5 ', t S5 is the temperature of satu- 
rated steam, r/ is the heat of vaporization, and q/ is the heat of 
the liquid, — all at the pressure p/; q is at the temperature t . It 
will be assumed also that t is less than t S5 and that the latter is 

* The " partial" pressure of the steam at the temperature T/ can be calculated 
approximately by the formula 

47 m 



p; = p 3 



29.3 + 47 »» 



if we assume the constants for the exhaust gases are the same as for air. In the 
same way the "partial" pressure of the steam after expansion is calculated thus: 

47 m 



29-3 + 47 m 



448 THE STEAM TURBINE 

less than t 5 ', as is generally the case, neglecting the small quan- 
tity of heat in the water vapor remaining in the mixture at the 
temperature t after the burned gases and steam have been dis- 
charged from the nozzle. 

We can write the following equations, applying' the same 
methods as for the case without the use of injection water: 

w/ =(<?,- Q 2 )y- ^-°> 

Z = • 

Qi 

Total efficiency, z, increases with the pressure of compression to 
a certain limiting value and then decreases. But for the practi- 
cable values of compressor and turbine efficiencies (x and y) the 
values of the theoretical total efficiency are not particularly good. 

The equations given here for velocity and efficiency can be 
used to investigate the best operating conditions by varying the 
pressure of compressions and the quantity of injection water. 

Another method for reducing the temperature of the gases is 
to use a large excess of air above the quantity needed to support 
combustion. The most economical method is probably that of 
partly vaporizing the CGoling water in the water jackets. The 
advantage of using water is that likewise it does not need to be 
compressed, and therefore it can be injected into the combustion 
chamber without the expenditure of much energy. 

Many of the troubles in the combustion chamber of a gas 
turbine are difficult to explain. One of the most serious diffi- 
culties is the occasional missnre of the incoming charge which is 
soon followed by a violent explosion. It is also difficult to secure 
smokeless combustion. Improvements are being made, how- 
ever, with the object of maintaining higher temperatures in the 
combustion chamber, and the results are encouraging. For this 
reason it is important that cooling water should be injected into 



GAS TURBINES 449 

the gases after they have left the chamber. The use of carbo- 
rundum for lining the combustion chamber and for the nozzles is 
apparently an important step forward. This material, which is 
a product of the electric furnace, is therefore manufactured at a 
much higher temperature than is ever attained in a gas com- 
bustion chamber. 



CHAPTER XVII. 
ELECTRIC GENERATORS FOR STEAM TURBINES. 

In the early years of the development of steam turbines it was 
the primary aim of the turbine engineer to reduce the speed of 
the turbine to operate satisfactorily when direct connected to 
electric generators. To accomplish this purpose De Laval 
introduced his famous helical gearing and Curtis applied the 
principle of velocity stages. To-day, however, the trend of 
developments is in the other direction. The electrical engineer 
is being urged to use his best skill to design generators to operate 
satisfactorily at higher and higher speeds, because in this way 
the efficiency of the turbine can best be increased. Very great 
strides have been made in the perfection of alternators for steam 
turbines ; but in the design of direct-current generators to operate 
at high speeds much is still to be desired. In fact, for high 
speeds, commutation is indeed a very difficult problem. The 
potential difference between adjacent commutator bars of engine- 
driven direct-current generators is usually about 10 volts; while 
for turbine-generators the limit is about 30 to 40 volts per bar. 

DIRECT-CURRENT GENERATORS. 

Sparking Limit. In slow-speed direct-current generators the 
output is limited either by the sparking or by the heating. On 
account of the necessarily small dimensions of the armature and 
the extremely high periodicity, with the resulting large iron losses, 
it becomes necessary, on the other hand, in a high-speed generator 
to employ artificial cooling devices. With the use of forced 
ventilation heating is no longer a factor limiting the output, and 
the difficulty lies then principally in the sparking. The quality 
of the commutation in any electric generator depends largely on 

450 



ELECTRIC GENERATORS FOR STEAM TURBINES 451 

the number of ampere-turns which can be placed on the surface 
of the armature; but there are also a number of other electric 
and magnetic conditions to be considered, particularly the effec- 
tiveness of the commutating poles,* located between the main 
poles. Another important factor is the mechanical condition 
of the armature, commutator, and brushes, which determines in 
a large measure the sparking limit. 

On the basis that the maximum permissible ampere-turns per 
centimeter of the circumference of the armature determine the 
maximum output the following table has been calculated. At 
an assumed permissible peripheral velocity of 75 meters f per 
second J (about 245 feet per second), this table shows the maxi- 



Revolutions 


Diameter of 


Output 


per 


Armature in 


in 


Minute. 


Centimeters. 


Kilowatts. 


4780 


30 


148 


2870 


5° 


347 


1800 


80 


670 


1435 


IOO 


890 


895 


160 


15% 


720 


200 


2080 



* When artificial commutation is secured by auxiliary poles placed between the 
main poles of a generator, short-circuit currents and sparking can occur only when 
the electromotive force induced by the commutating field is different from the 
reactance voltage 

t The C. G. S. (metric) system of units is applied in this chapter because it is 
the one commonly used by designers of electrical machinery in America and 
in England. 

% It is usually found that the end shells or shields protecting the connections in 
a revolving armature have stresses most nearly approaching the allowable limits. 
Stresses in these end shells are calculated as in a ring or band, by the following 
equation, as expressed approximately in C. G. S. units, 

Sa 



where V is the peripheral velocity of the ring in meters per second, 2 is the weight 
of a cubic centimeter of the material, and Sa is the allowable unit stress in kilograms 
per square centimeter. Since the allowable permissible stress of bronze castings is 
about 260 kilograms per square centimeter, the maximum allowable velocity is only 



452 THE STEAM TURBINE 

mum outputs and speeds which at the present time are obtain- 
able in the very best designs of direct-current generators.* 

This table shows that the armature of direct-current generators 
cannot be constructed to give the required output at the usual 
speeds adopted in America for Parsons turbines, and that with 
Curtis turbines, which operate at slower speeds, the limit is 
reached at 1500 kilowatts full load capacity. 

There are two ways of overcoming the limitations of direct- 
current generators for turbine service. One way is to design the 
turbines for lower speeds, which entails, however, increased cost 
and a sacrifice of economy. The other way is to adopt the tan- 
dem arrangement of connecting two generators to one turbine 
as in the usual De Laval designs for the larger sizes. 

There is a constant demand for direct-current generators of 
larger capacities than are now employed, and the problem of 
increasing the capacity of the generator is becoming very impor- 
tant. The successful production of such machines suitable for 
much higher speeds than are now attainable would be an improve- 
ment effective in two ways: (1) by lowering the steam consump- 
tion and (2) by reducing the first cost; and as a result the field 
of the high-speed reciprocating engine would be still more 
restricted. 

Flash-over Limit. In ordinary slow-speed direct-current 
generators the only electrical limit to the capacity is sparking. 
In high-speed machines, however, a new difficulty known as the 
flash-over limit is met. Its effects are often as serious and as 
difficult to remedy as any of the commutation troubles. A great 
many designs of high-speed direct-current generators with 
satisfactory commutating qualities have been failures because 
of their tendency to arc around the whole commutator. This 
trouble must be attributed primarily to the very high potential 

55 meters per second; and considering the additional load due to the end connec- 
tions the permissible velocity becomes only about 50 meters per second. If, how- 
ever, phosphor-bronze or manganese-bronze castings with an allowable stress of 
600 kilograms per square centimeter are used, a peripheral velocity of 75 meters 
per second is not excessive. 

* R. Pohl, Proc. of Inst, of Elec. Engrs., 1907. 



ELECTRIC GENERATORS FOR STEAM TURBINES 453 

difference between adjacent commutator bars — usually about 
three times the permissible value in slow-speed generators. 
Usually this difficulty can be remedied by increasing the insula- 
tion of the shrinkages and of the brush-gear. It is the flash- 
over limit, therefore, which determines the allowable voltage 
per commutator bar and restricts the number of " lines" or flux 
allowed to enter or leave an armature of a given diameter. 

The most obvious line of improvement in turbine-driven 
direct-current generators is in increasing the peripheral speed of 
the armature. Steel alloys of very low magnetic conductivity 
and high tensile strength used in the place of phosphor-bronze 
for the end shields of the armature will permit the adoption of 
considerably higher peripheral speeds than are now allowable. 
If we compare two machines of equal output and speed but with 
armatures of different diameters in the ratio of 1 to 2, the arma- 
ture with the larger diameter will be only one-fourth as long as 
the other; while with the same voltage for both the number of 
conductors and segments will be doubled.* 

. ALTERNATING-CURRENT GENERATORS. 

The design of alternators with revolving fields to operate at 
high speeds is not nearly so difficult as for commutating machines. 
Speed limits are usually determined by the strength of suitable 
materials for their construction. When it became the general 
practice to enclose high-speed generators in a sheet-metal casing 
and to adopt forced or artificial ventilation produced by small 
fans circulating air through the generator windings, it was 
possible to regulate the heating limit so that heavier overloads 
could be carried and for longer periods of time than was possible 
before. By this method the excessive noise of the early turbine- 
generators was, at the same time, eliminated. 

For the windings of the revolving fields of high-speed alterna- 
tors, flat strap copper is used by most manufacturers. To make 
the field spools as small and compact as possible this strap copper 

* The voltage per segment is approximately inversely proportional to the 
peripheral velocity of the armature. 



454 



THE STEAM TURBINE 



is sometimes coated with a thin layer of enamel for insulation, 
instead of the usual cotton covering. The necessity of making 
the exterior surfaces of revolving fields as smooth as possible is 
generally appreciated by designers. 



GENERATOR EFFICIENCIES. 



Average efficiencies of the best designs of alternating-current 
generators intended for operation with steam turbines are given 
in the following table: 



Rated Full Load Capacity, 
Kilowatts. 


Efficiency of 
Alternator, 
Per Cent. 


5° to 1 5° 
200 tO 4OO 

500 to 900 
1000 to 2500 
3000 to 5000 
6000 to 10,000 


90-93 
93-94 

95 
96 

97 
98 



The efficiency of direct-current high-speed generators is about 
one per cent, less than that of alternators of the same capacity. 



APPENDIX. 
EXERCISES ON STEAM TURBINES. 

Exercise i. What is the velocity of steam discharging at the rate of 200 
cubic feet per second through a nozzle having a cross-sectional area of 0.2 
square foot? Ans. 1000 feet per second. 

Exercise 2. If the steam discharging from the orifice mentioned in the 
preceding exercise weighs .0322 pound per cubic foot, how much energy 
in foot-pounds per second can this jet develop ? How much horsepower ? 
Ans. 100,000 foot-pounds per second. 181. 8 horsepower. 

Suggestion: From elementary mechanics we have the information that 
the kinetic energy K (sometimes called capacity to do work) of any moving 
fluid, such as steam, gas, or water, is 

.&. = 1 

2g 

where W is the weight of the fluid discharging per second, V is the velocity 
of flow in feet per second, and g is the acceleration of gravity or 32.2 feet 
per second. 

By definition (in English units) one horsepower is equivalent to 550 
foot-pounds per second. 

Exercise 3. If the vessel shown in Fig. 33 discharges 40 pounds of water 
per second at a velocity of 161 feet per second, what is the force (impulse) 
pushing the wooden block away from the vessel? Ans. 200 pounds. 

Also what is the force (reaction) pushing the vessel itself toward the 
left? Ans. 200 lbs. 

Exercise 4. If water is discharged against flat blades of a water wheel 
made up of vanes similar to the block shown in Fig. 33 (page 60) at the 
rate of 3.22 pounds per second at a velocity of 2000 feet per second and is 
spattered from the wooden blocks with a "residual" velocity (leaving the 
vanes) of 300 feet per second, what horsepower is this water wheel capable 
of developing? Ans. 195,500 foot-pounds per second or 355.5 horsepower. 

Exercise 5. Steam of the same density as in the exercise at the bottom 
of page 66 discharges at the rate of 1739 pounds per hour and produces a 
reaction against the plate into which the same nozzle is inserted of 45 pounds. 
What is the velocity of discharge ? Ans. 3000 feet per second. 

455 



456 APPENDIX 

Exercise 6. The area of a nozzle at its smallest section is .72 square inch 
and discharges steam at the rate of .2 pound per second, of which the specific 
volume is 2.0 cubic feet per pound. 

(a) What is the velocity of flow? ' Ans. 80 feet per second. 

(b) What is the magnitude of the force developed by the reaction of 
this jet? Ans. | pound (nearly). 

(c) What is the maximum value of the impulse produced by this jet if 
friction and eddy losses reduce the velocity effective for giving the impulse 
by 25 per cent.? Ans. .37 pound. 

(d) If only part of the velocity available in (c) is absorbed in driving a 
steam turbine, so that the steam leaves the blades with a "residual" veloc- 
ity of 10 feet per second, how many foot-pounds of work per minute are 
developed by the turbine ? Ans. 652. 

(e) What is the horsepower equivalent of this number of foot-pounds 
per minute? Ans. .0198 horsepower. 

(/) If this turbine drives a small electric generator having an efficiency 
of 80 per cent., what power in kilowatts will this generator develop? 

Ans. .0119 kilowatt. 

Suggestion: A kilowatt is a thousand watts, and 746 watts are equiva- 
lent to a horsepower. 

(g) How much horsepower would be developed by this turbine if all the 
velocity as calculated in (a) is transformed into work ? 

(h) What is the efficiency of the turbine ? 

Suggestion: Compare (e) and (g). If the velocity as calculated in (a) 
represents the total velocity equivalent of the available energy due to 
adiabatic expansion (constant entropy) then the answer to section Qi) is 
called the Rankine efficiency of the turbine. 

Exercise 7. Calculate the horsepower developed by a steam turbine 
having two rows of moving blades. Upon the first row steam is directed 
at a velocity of 3000 feet per second and at the rate of 1.7 71 pounds per 
second. The steam is discharged from this row at a velocity of 1000 feet 
per second and is then directed upon a second row of blades from which it 
is discharged at a velocity of 200 feet per second. 

(a) Neglecting frictional and other losses, how much horsepower will 
this turbine develop ? Ans. 448 horsepower. 

(b) How much power would be developed if there is a loss of velocity of 
10 per cent, in each row of blades ? 

Suggestion: Actual velocity effective in first row of blades is 2700 feet 
per second which is discharged at 1000 feet per second. Work developed 
in this row is 1.771 (2700 2 — iooo 2 ) -5- (64.4 X 550) in horsepower. Simi- 
larly work done in the second row is 1.77 1 (900 2 — 200 2 ) 4- (64.4 X 550) 
in horsepower. 



APPENDIX 457 

(c) What is the efficiency of the complete turbine when the losses stated 
are considered? 

Suggestion: Efficiency is total horsepower calculated in (b) divided by 
that found by considering only initial velocity (3000 feet per second) as 
in (g) of Exercise 6. 

Exercise 8. (Use of entropy- total heat chart.) 

Steam at an initial condition of 165 pounds per square inch absolute 
and 100 degrees Fahrenheit superheat is expanded in a nozzle adiabatically 
to 20 pounds per square inch absolute pressure. 

(a) How much energy (in B.T.U. per pound) is converted into velocity? 

(b) If no losses are considered, what is the velocity in feet per second of 
the discharging jet ? 

(c) If there are losses equivalent to 4 per cent, of the energy available, 
what is the actual velocity of the jet ? 

(d) If the losses are equivalent to 2 per cent, of the theoretical velocity, 
what is the actual velocity of the jet ? 

Exercise 9. Steam at the same initial and final conditions as in Exer- 
cise 8 is reheated by friction in the nozzles and blades so that the entropy 
at the final condition is 1.7. How much energy is available for doing work? 

Suggestion: Reading from the entropy-total heat chart, the total heat 
contents of a pound of steam at the initial condition is 1252 B.T.U. and at 
the final condition after reheating is 113 5 B.T.U. (expansion is not adia- 
batic) . Therefore, heat units available for work =1252 — 1 13 5 or 117 B.T.U. 

Exercise 10. A certain steam turbine having several stages takes steam 
initially at 165 pounds per square inch absolute and 100 degrees F. 
superheat and expands it to 20 pounds per square inch absolute in the 
first stage. Friction and the transformation of residual velocity into 
potential (heat) energy returns 30 per cent, of the available energy in an 
adiabatic expansion back to the steam by "reheating" at the final pressure. 
If now in the nozzles of a succeeding stage of the turbine the steam is 
expanded to 10 pounds per square inch absolute and reheated agajn by the 
same percentage at the latter pressure, how much energy (B.T.U.) is avail- 
able in each stage for performing work? What is the quality of the steam 
after each reheating? 

Suggestion: In the normal adiabatic expansion from 165 pounds per 
square inch absolute and 100 degrees F. superheat to 20 pounds per square 
inch absolute the available energy is 1252 — 1085 or 167 B.T.U., of which 
30 per cent, or 50.1 B.T.U. go to reheat the steam. The total heat con- 
tents of a pound of steam after reheating becomes then 1085 -f- 50.1 or 
1 135. 1 B.T.U. For the second stage the expansion is from 20 pounds per 
square inch absolute pressure and a total heat contents of 1135.1 B.T.U. 
(quality about .978) to 10 pounds per square inch absolute, making the 



458 APPENDIX 

available energy for adiabatic expansion 1135.1 — 1087 or 48.1 B.T.U. 
per pound. The reheating is 30 per cent, of this or about 14.3 B.T.U., which 
when added to 1087 gives 1091.3 B.T.U. as the total heat contents of the 
steam when passing into the nozzles of the next succeeding stage. At this 
condition the quality of the steam is about .949. 

In actual designing the reheating in the last stage is not considered 
available, as will be observed in the design worked out on page 103. The 
reason for this is that a very large part of the reheating in the stages other 
than the last is due to the changing of the residual velocity of the steam 
as it leaves the blades into potential (heat) energy. This has been dem- 
onstrated by actual experiments which show that the steam enters the 
nozzle of an impulse wheel in every stage with practically negligible velocity. 
In the last stage the conditions, however, are different. The steam here 
leaves the blade with its residual velocity unchecked and passes off into 
the large exhaust passages provided for its unimpeded flow. 

Exercise 11. A turbine blade like the one shown in Fig. 43 moves with 
a velocity of 500 feet per second due to a steam jet passing over it which 
has a velocity of 3220 feet per second. If friction losses in the blade are 
not considered, and the weight of steam flowing per second is 1.0642 pounds, 
a is 20 degrees and = 7 = 45 degrees, what is the total impulse force 
to which the blade is subjected in the direction of its motion (see page 76)? 

Suggestion: Since losses in the blades are neglected F r3 = V r2 and V r2 = 
VFfc 2 + V 2 2 - 2 V b V 2 cos a. (Law of Cosines). 

All the terms in this equation are known so that V r2 or V r3 can be calcu- 
lated. 

Exercise 12. Taking the necessary data from the preceding exercise, 
state the proper angle for the backs of the blades (see Figs. 49 and 50), 
for the steam to enter without loss due to impact and eddying. (See page 
75.) 

Exercise 13. Explain the essential principle of operation of Hero's 
engine. Indicate clearly in a figure the direction of rotation of this engine 
with respect to the direction of steam discharge from the nozzles. What 
is the difference in principle between Hero's engine and Branca's? 

Exercise 14. Explain the actual difference between the commercial types 
known as impulse and reaction turbines. 

Exercise 15. Why have the stationary blades or buckets shown in 
figures like 39, page .68, a curvature in the opposite direction to that of all 
the moving blades ? 

Exercise 16. Why is the rotation loss when stated in per cent, of rated 
output less for a large size turbine than for a relatively small one? 

Exercise 17. Design a nozzle, showing all the important dimensions, 
for expanding steam from the initial condition of 165 pounds per square 



APPENDIX 459 

inch absolute, and ioo degrees F. superheat to a final condition of 4 pounds 
per square inch absolute. Assume that the nozzle loss is 3 per cent, of the 
velocity and that the rate of flow is to be ts pound per second. (See pages 
ioo-in.) 

Exercise 18. Dry steam expands in the nozzles of a simple impulse tur- 
bine from 165 pounds per square inch absolute to 1 pound per square inch 
absolute (about 28 inches vacuum). Draw velocity diagrams, allowing for 
no losses, and determine the proper blade angles when p equals y. The 
nozzle angle is to be, as usual, 20 degrees and the peripheral speed of the 
blades or buckets is 1200 feet per second. 

Calculate the energy absorbed or given up to the blades or buckets per 
pound of steam as well as the steam consumption of the ideal turbine 
(theoretical water rate) and the steam consumption of this turbine as de- 
termined from the energy absorbed by the blades or buckets. 

Sketch with a reasonable degree of accuracy the outlines of the blades or 
buckets. 

Exercise 19. Recalculate and redesign the blades for the conditions 
given in Exercise 18 when the nozzle loss is 3 per cent, of the theoretical 
velocity developed and the blade losses are obtained from Fig. 51, page 101. 

Observe and discuss the change in blade angles caused by including the 
losses in the design. 

Calculate (1) the work done in foot-pounds per second per pound of 
steam; (2) the steam consumption per horsepower-hour and the efficiency 
of the turbine. 

If the speed of the turbine is 20,000 revolutions per minute, find the 
diameter of the mean blade circle. 

If five nozzles are used for a maximum load of 50 horsepower, find the 
diameter at the throat of each of these nozzles, assuming they are all of the 
same size. 

Exercise 20. Make the necessary calculations and draw velocity diagrams 
and neat sketches of the blades for an impulse turbine having two pressure 
stages and two rows of moving blades, that is, two velocity stages in each 
pressure stage, for the following requirements: 

The initial pressure of the steam supplied to the turbine is 165 pounds 
per square inch absolute and is expanded in the first set of nozzles to 20 
pounds per square inch absolute. In the second set of nozzles the pres- 
sure falls from 20 pounds per square inch absolute to 2 pounds per square 
inch absolute (about 26 inches vacuum). The nozzle angles are 20 degrees 
and the peripheral speed of the blades or bucketsjs 500 feet per second, the 
nozzle loss is 2 per cent, of the theoretical velocity, and the blade losses are 
to be taken from Fig. 51. Assume that the windage, leakage, and bearing 
losses amount to 30 per cent, of the energy developed by the action of the 
steam in the blades. Steam is initially superheated 100 degrees Fahrenheit. 



460 APPENDIX - 

The rating of the turbine is to be for ioo horsepower at 1800 r.p.m. 
Calculate the number of buckets and the height of the buckets for the 
first row in the first stage and for the last row in the second stage. 

Observe that the height of the blades for the first row in each stage is 
determined by the height of the nozzles which discharge into the blades. 
\ Exercise 21. Design the blading of a reaction turbine for the same con- 
ditions given in the second exercise on page 145, except that the initial 
steam pressure is to be 175 pounds per square inch absolute, and the final 
pressure § pound per square inch absolute. 

Exercise 22. Design a combined impulse and reaction turbine, taking 
the general data the same as for the preceding exercise and the expansion 
in the impulse section to be from 175 pounds per square inch absolute to 
40 pounds per square inch absolute. The expansion in the reaction blading 
is to be from 40 pounds per square inch absolute to £ pound per square inch 
absolute. 

Sketch the blades for the impulse section, assuming there are two velocity 
stages, and also the blades for the first and last stages in the "reaction 
section." 

Exercise 23. Determine the velocity loss in feet per second in a nozzle 
having 98 per cent, efficiency at its proper expansion, which is from 125 
pounds per square inch absolute pressure to 28 inches vacuum (referred to 
30 inches barometer) when used for 

(1) 165 pounds per square inch absolute and 29 inches vacuum. 
I (2) 9 pounds per square inch absolute and 26 inches vacuum. 

State also the corresponding energy loss in B.T.U. per pound of steam 
in each case, and by what percentage the efficiency of the Rankine cycle 
will be affected. By what percentage would the steam consumption of a 
commercial type of turbine be affected? In all cases mentioned the steam 
is initially dry saturated. 



INDEX 



Absolute velocity 74 

Accumulator, Rateau 327-330 

Adiabatic expansion 21-24 

Adjustment bearing IQ 8, 217 

Allgemeine Electrizitats Gesellschaft Turbines (Curtis and Riedler- 

Stumpf) 230, 259 

Allis-Chalmers-Parsons turbines 209-212 

blades 203-205 

governor 294 

rotor 41 2 

shroud rings 203 

Alternators, parallel operation 391 

with forced ventilation 453 

Area of floor for engines and turbines 391-396 

nozzles 37~49 

Auxiliary machinery, power for 388 

Available energy ; 21-26 

Axial clearance of blades 147 

Balance pistons 136, 198, 217 

Bearing friction 108, 161, 186 

Blades, clearances of 127, 145, 147 

conditions of best efficiency 75-87 

design of 97-141 

erosion of 150 

height of 98, in, 135, 137, 143 

impulse and reaction upon 60-64 

radial leakage of 27, 131, 135, 137, 147, 161 

materials for 147-150 

peripheral speeds of 102, 124, 133, 134, 180 

pitch of 98, 130 

rotation loss of 153 

Bleeder steam turbines 339 

Branca's turbine 5 

British-Westinghouse turbine 209 

British thermal unit 12 

Brown-Boveri-Parsons turbine 209, 292, 305 

Buckets (see also Blades) 230, 259, 269 

By-pass governing 304-309 

Centrifugal force 4°5 

Chart, entropy — total heat (in Appendix). 

461 



462 INDEX 

PAGES 

Clearances of blades, axial 147 

radial '. 127, 145, 147 

Commercial testing of turbines 355 

Comparison of turbines and engines 162-167, 173 

Composition of blades 147-150 

Condensing water, quantity of 372 

Condensers and auxiliaries, cost of 385-388 

power for 389 

Corrections for economy curves 162-172, 379 

for rotation losses 154 

for steam turbine tests 162-172 

Cost of engines, turbines, and auxiliaries 383-388 

maintenance and operation 382-383 

Critical speed of rotating shafts 8, 183, 430 

Curtis turbines 230-243, 340 

analysis of losses 244 

blades 149, 233 

diaphragms 230 

emergency stop valve 238 

governor 239, 284, 288 

manufacturers 230 

nozzles 27, 232 

oiling system 4 00- 4°3 

small sizes 240 

speeds 238 

steam consumption 169, 243, 358 

step bearing 235 

superheat corrections 169, 242 

tests on 169-172, 358 

vacuum corrections 169, 242, 370 

valve gear, electric type 282 

hydraulic 285 

mechanically operated 282-285 

wheels 232, 234 

Dake turbine 269 

De Laval turbine 5~ 6 > 176-193 

analysis of losses I 94 

bearings l8 3 

blades 149, 182 

disks z 7% 

gears l8 4 

governor 186, 276, 302 

nozzles 3> 2 7> 180 

shaft l82 

speeds of l8 ° 

steam consumption x 94 

superheat corrections x 9 2 



INDEX 463 

PAGES 

De Laval turbine — Continued 

tests on, 362 

vacuum corrections 193 

wheels 178 

Depreciation of power plants 386-387 

Design of power stations 389-399 

Design of turbines, examples in 102-145 

entropy-heat, diagram applied to 103 

of blades 97-101 

of steam nozzles 3i - 59 

Deterioration of turbines, and auxiliaries 386-387 

Diaphragms .... . 97, 230, 248 

Disk rotation losses 153 

stresses in 414-430 

Drums, stresses in... 407-412 

Economy of engines, best. ... 380 

compared with turbines 377 - 38o 

with varying superheat, vacuum, and pressure 381 

Economy of small engines and turbines 381-382 

Economy of standard turbines 379 

Efficiency of blades 101 

steam engines, mechanical 378 

steam nozzles 58, 102 

thermal unit basis of . 562-^6^ 

turbines 101-102, 362-363 

turbine-generators 454 

Electric generators for steam turbines 450-454 

End thrust , 118, 198 

Entropy 17 

of saturated steam 18-23 

of superheated steam 19, 26-31 

of water 19-20 

temperature diagrams 17-25 

total heat chart 43, 45, 47, Appendix 

Erosion of blades 150 

Experiments with nozzles 49~58 

on flow of superheated steam 39~4* 

with governors 3°9 

Extruded metal for blades 148 

Flash-over limit of turbine-generators 45 2 

Floor area for engines and turbines 391-396 

Flow of steam: 

experiments on 36-42 

Grashof's law for 3°-39 

Napier's formula for 36 

saturated 36-39 

superheated 39"4i 



464 INDEX 

PAGES 

Flow of steam — Continued 

weight of steam flowing 36-41 

Gas turbines \ 432-449 

compared with gas engines 441 

Gauging of Parsons blades 118, 126, 134 

Generators, efficiency of 454 

Glands, water packed 200, 255 

Governing of turbines, methods of 274-310 

by-pass 304-309 

cutting out nozzles t 277-285 

experimental data 309 

throttling 274-277 

varying time of admission 285, 292-304 

Governors, turbine: 

Allis-Chalmers 294 

Brown-Boveri 292, 305 

Curtis 282-285, 288 

De Laval 276-277, 302 

Westinghouse 298-300 

Wilkinson 302-304 

Grashof's law for flow of steam 36-39 

Hamilton-Holzwarth turbine 256 

Heat diagram, entropy Appendix 

applied to design 103 

Heat theory 10-28 

Heat units 12 

Heat, mechanical equivalent of 13 

specific 1 2, 13, 26-29 

Hero's turbine 4, 8 

Horse-power, internal or indicated 364 

conversion table for kilowatts to brake 378 

Impulse turbines 62, 67, 176 

and reaction, combined (Westinghouse) 215-229 

blade design 97~i45 

blade efficiency 75-87 

blade losses 100-102 

comparison with reaction turbines 145 

distinguished from reaction 60-64 

shape of blades 63-64 

Initial velocity of steam 16 

Injection water (see Condensing water). 

Intermediate (stationary) blades 69, 233 

Jets, impulse and reaction of 60-62 

Kerr turbine 263, 266 

governor 265 

Kilowatts, conversion of, to brake horse-power 109 (footnote), 378 

Knoblauch and Jakob's specific heat of superheated steam 27-29 



INDEX 465 

PAGES 

Labyrinth packing 198, 209, 255 

Lasche's method for rotation losses 158-160 

Leakage of steam through blades 127, 131, 135, 137, 161 

through diaphragms 101, 161, 255 

Losses in a turbine 101, 108, 151-161, 194 

in a De Laval turbine, analysis 194 

Low-pressure turbines 311-334 

accumulators for 327-341 

combined with gas engines 334, 383 

steam consumption of 315, 330-333 

Lubrication of turbines 300, 396, 400-403 

cost of oil required 383 

of Curtis turbines 402 

of Westinghouse turbines 298-301 

Maintenance and operation of turbines 382-385 

Marine turbines 313-315, 345"352 

designing constants 117 

working with reciprocating engines 313 

Materials and permissible stresses in turbine disks 429 

Mechanical equivalent of heat 13 

losses in turbines 151-161 

Mixed pressure turbines 335~338 

Moisture correction for rotation losses 154 

Moisture correction for steam consumption (water-rate) 39, 171 

Monell metal 150 

Monnot metal 148 

Motion, absolute and relative 74~75 

Napier's formula for flow of steam 36 

Non-expanding nozzles 55, 59, 1 73, footnote 

Nozzles : 

area of 36-45 

design of 36-59, 1 10 

efficiency 57 _ 58 

examples of 33, 34, 59 

expansion ratio 44-50 

flow of steam through 36-59 

length of 54 

non-expanding 55, 59, 215, 248 

types of 3, 33, 34, 59 

losses in 57—58 

Oil pump, Westinghouse 298-301, 401 

Oiling systems 298, 396, 400 

Curtis 402 

Over- and under-expansion in nozzles 57-58 

Packing glands 191, 200, 255 

Parsons turbines 6-9, 195-214 

bearings 8, 206 



466 INDEX 

PAGES 

Parsons turbines — Continued 

blades and lashing . 200-205 

design of 111-145 

governors and valve gears .... 1 -. >4 , » 1 » 285, 292, 294 

history •. . , , , , 6-9 

manufacturers of 209 

number of stages 114-117 207 

pressure diagram (indicator) . . . -.-.. 2 qj- 

shroud rings 201-203 

speeds of II4 

steam consumption of .,„ 169-172 212 

superheat corrections 169, 214 

vacuum corrections -...-. ; 169, 214 

Pelton types of turbines 259-269 

Performance, thermal unit basis of 362-363 

of engines and turbines joi, 378 

Piping of turbine stations 397~399 

for superheated steam 403 

Power plant economics 365, 382 

Pressure (initial steam) effect on economy -. 376 

corrections for economy curves 162-172 

in throat of nozzles ^6 

stages 70 

Pressure-volume diagrams rS-17 

Prices of turbines and auxiliaries 383-385 

Quality of steam 39, 44 

Radial blade leakage 27, 131, 135, 137, 147, 161 

Rankine cycle efficiency 31, 109, 363, 364 

Rateau accumulator 327-330 

Rateau turbines 247 

diaphragms 248 

low pressure 252 

manufacturers 252 

tests on 331 

Reaction of jets 60-62 

Reaction turbine 73 

blade design 111-121 

blade efficiency 84-87 

blade losses 100-101 

distinguished from impulse 63-64 

Reheating factor 102, 103, 123 (footnote) 

Relative velocity 74~75 

Riedler-Stumpf turbines 259 

Rings, stresses in 406 

Rosenhain's tests on flow of steam 55 

Rotation losses 77 (footnote), 151 

factors to correct for superheat and moisture 154 



INDEX 467 

PAGES 

Rotors, stresses in 409-41 2 

Sankey's blading 203 

Screw turbine 9 

Shroud rings 203, 233 

stresses in 406-408 

Space occupied by engines and turbines 391-396 

Sparking limit of generators ...... 450 

Specific heat 12 

of superheated steam 13 

Specific volume of superheated steam 40, 122 

Speed of turbines, effect on economy and output 174, 376 

Stages of turbines 7° _ 7 I 

leakage between 83 (footnote), 101, 161, 255 

Steam turbine economics 365-404 

Steam turbines: 

compared with steam engines 1, 13-17 

cost of 383-386 

Step-bearing of Curtis turbines 235 

Stodola's experiments with nozzles 56-58 

Stresses in rotating rings, drums, and disks 405-431 

at right angles 4 1 2 

Sturtevant turbine 259-263 

bearings 262 

buckets 259 

nozzles 260 

velocity stages 262 

wheels 263 

Superheated steam corrections for economy curves 162-172, 379 

effect on economy 373*376 

flow through nozzles 39~4° 

specific heat of 26-30 

specific volume of * 40, 122 

total heat of 28-29 

Taper of De Laval nozzles 54 

Temperature-entropy diagrams i7 _2 9 

Temperature 10-1 1 

absolute 1 1 

Terry turbine 266 

Tests on turbines 168, 353-363 

Thermal units, definition of 12 

basis of performance 362 

Thermodynamic efficiency 363, 364 

Thermometer correction n 

Thrust, end, in Parsons turbine 118, 198 

Torque line 1 74 

Total heat of superheated steam 28-29 

Types of steam turbines 62-64, 176 



4 6S INDEX 

PAGES 

Vacuum, condensing water required for 367, 372 

corrections for economy curves 162-172, 370 

effect on engine and turbine economy 365, 380 

most profitable 365 

Vanes (see Blades). 

Velocity, absolute 74-76 

diagrams 68-74 

relative 74, 75 

Velocity of blades 102, 124, 133, 180 

Velocity of steam: 

calculation of 24-26 

stages 70, 82, 215 

Volume, specific: 

at high vacuums (table) 312, 314, 369 

of superheated steam 40, 122 

Water-packed glands 200 

Water-rate curves 31, 354 

Westinghouse oil pump 401 

Westinghouse impulse turbines 243-245 

Westinghouse impulse and reaction turbines 215-223, 246 

adjustment bearing 217 

advantages of 223 

blades of impulse section 215 

emergency speed limit 219 

nozzles and nozzle block 219, 220 

reduction gear 347 - 35o 

velocity stages 215 

Westinghouse-Parsons turbines 207 

bearings 206 

blades and lashing 201, 202 

governor and valve gear 293, 298 

tests on 169, 213, 363, 364 

water-packed glands 200 

Wilkinson turbines 255 

governor and valve gear 302 

stage packing 255 

Willans and Robinson turbine 203 

Willans lines 160 

Wing blades <)C 119, 129 

Zoelly turbines ^0. 257 

blades of .^j^ 257 

governor ^L- 3°9 

nozzles. \ff. 257 

tests on Jte 362 

wheels ;. 258 > 



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